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THE GASOLINE AUTOMOBILE 


Its Design and Construction 


Volume I 


The Gasoline Motor 



By 

P. M. HELDT 

4 * 

Editor of The Horseless Age 


THIRD EDITION, REVISED 



PUBLISHED BY 

THE HORSELESS AGE CO. 

250 West 54th Street 
NEW YORK 
1915 


xX-^4 


o 

N 


Copyrighted by 

THE HORSELESS AGE COMPANY, 
1911, 1912 and 1915. 


(All Rights Reserved.) 


MAY 11 1915 

©CI.A397949 


/ / 


<flS 




PREFACE TO FIRST EDITION. 


The matter which appears here in book form was originally 
published serially in The Horseless Age. This method of pub¬ 
lishing a work possesses the advantage that it permits of offering 
the book at a lower price than would otherwise be possible. It 
is believed that many of those who read the articles in the paper 
will also buy the book, on account of its more handy form; their 
familiarity with the contents will enable them to use the book to 
greater advantage as a work of reference. Most engineers have 
probably had the experience that, aside from the so-called engi¬ 
neers' pocket books, the most useful books are the text books 
used during their college courses, and this because they know 
just what to look for in them, and where to find it. 

Although the automobile industry has developed in recent years 
at such a rapid rate that it now ranks among the leading indus¬ 
tries of the country, there has not been published a handbook 
dealing comprehensively with the subject of automobile design. 
It is true that two or three small volumes on automobile engineer¬ 
ing are extant, but they are quite limited in scope and leave unan¬ 
swered many questions which come up in designing a motor car. 
The need for a comprehensive work on the subject has often been 
strongly impressed upon the author, who has frequently been 
asked to name or recommend a book of this nature, and not 
being able to do so, he resolved to make an attempt to supply 
this deficiency in automobile literature. 

It is proposed to cover practically all of the questions with 
which the automobile engineer is directly concerned, in two 
volumes. The first volume, now completed, deals with the gaso- 

III 


IV 


PREFACE. 


line motor, and the second will treat of the running gear and the 
transmission. Ignition appliances, carburetors and radiators will not 
be dealt with in either of these volumes, since these apparatuses 
are now the object of special industries, and are very rarely built 
or designed in the automobile factory. Body construction is also 
a subject apart. 

In writing this first volume the author endeavored to confine 
himself mainly to an exposition of the principles underlying the 
design of the gasoline motor, and closely to limit historical notes 
and descriptions of individual designs. Detailed technical de¬ 
scriptions were considered unnecessary, because there is no dearth 
of literature of that nature. The automobile periodicals almost 
every week publish descriptions of new motors, and any one de¬ 
sirous of familiarizing himself with the individual features of 
different makes can do so by perusing these periodicals. It was, 
however, found impractical to entirely exclude descriptive matter 
and historical allusions. 

The phenomena occurring in the cylinder of a high speed gaso¬ 
line motor still remain more or less obscure. The invention of 
the manograph or optical indicator, about a decade ago, paved the 
way for the experimental investigation of these phenomena, but 
little has as yet been done in that line, at least so far as one may 
judge from the meagerness of published results. On the other 
hand, by continued experiment the gasoline motor has been 
brought to a high state of perfection. The operation of the motor 
is unquestionably based upon the established laws of physics and 
mechanics, yet it has been difficult to build up a working theory 
upon these fundamental laws and develop it to a point where 
it was possible to design a practical engine with its aid alone. 
The present work aims to bring theory and practice into closer 
relationship. Formulae are developed for the dimensions of all 
the more important parts of the engine, which generally involve 
the bore and stroke, the particular physical quality of the ma¬ 
terial used, and some dimension of the part which can be as¬ 
sumed. But all of these equations embody a constant, and the 
values of these constants were determined from the dimensions 
of the part in question on numerous modern engines regarding 


PREFACE. 


V 


which data was available. The value of the constant in any equa¬ 
tion was determined for each engine, and the average of the 
values thus found was then taken. The formulae given may thus 
be said to represent good average practice. Assembly drawings of 
engines, many of full size scale, were furnished the author by 
more than a score of well known manufacturers, and he wishes 
to take advantage of this opportunity to express his thanks for 
the courtesy. Many of the prints are reproduced in the folding 
plates. 

In a work of this nature it is, of course, impossible entirely to 
avoid the use of mathematics, since machine design consists 
chiefly in practical application of mathematics and mechanics in 
conjunction with experimental data. But it has been the aim 
to make the mathematical treatment as clear as possible, and in 
order to facilitate the use of the formulae derived, their applica¬ 
tion is illustrated by practical examples in most instances. 

As editor of The Horseless Age during the past eleven year* 
the author has had the good fortune to edit a large proportion 
of the automobile engineering literature that has been published 
in this country during that period. He has also been a constant 
reader of the more technical foreign automobile papers, and is 
fairly familiar with foreign automobile book literature. All of 
the above mentioned sources of information have been drawn 
upon to some extent. Where matter has been taken directly 
from any publication due credit is given. The chief source has 
been the pages of The Horseless Age, and where extracts are 
made from signed articles, credit is given to the writers thereof. 
It has not been deemed advisable to follow the prevailing prac¬ 
tice of giving numerous references to articles in engineering peri¬ 
odicals, because it has been the experience that such references are 
of little practical value to readers, with the possible exception 
of those located in the two or three largest cities in the country. 

Incorporated at the end of this volume are a considerable num¬ 
ber of folding plates showing part-sectional views of modern 
American and European motors of different types. 

No one will dispute the fact that the rapid progress of auto¬ 
mobile engineering has in a large measure been due to the lib- 


VI 


PREFACE. 


eral exchange of ideas and experimental results by engineers in 
the automobile periodical press. In fact, until now the technical 
papers largely served as sources of engineering data from which 
scrap books were made up. As long as automobile engineering 
was still in an unsettled state this was perhaps the best plan. 
Now, however, that things are settling down to a firm basis, 
there is need for a handbook in which information on any phase 
of the subject may be readily looked up by those engaged in the 
art. 

Another important fact is that as the industry grows the firms' 
engaged in it, and new ones entering it, are obliged constantly to 
put on new men in their engineering and draughting departments. 
Only comparatively few men have had the advantage of “grow¬ 
ing up with the industry”; and the others must necessarily fa¬ 
miliarize themselves with automobile engineering practice by 
reading, since no schools teach the subject as yet. But even if 
the technical schools should take the matter up there would still 
be need for a suitable text book. It is, then, with the object of 
meeting the requirements of both of these classes—those actively 
engaged in automobile engineering and those wishing to prepare 
themselves for an automobile engineering career—that this work 
was written, and it is hoped that they will find it satisfactory. 

P. M. Heldt. 

New York, November, 1911. 



\ 


LIST OF CHAPTERS. 


Page. 

Chapter I. The Working Media—Gasoline and 

Air, . i 

Chapter II. The Laws of Gases—Thermodynamics, io 

Chapter III. The Otto Cycle,.19 

Chapter IV. Conversion of Reciprocating Into Ro¬ 
tary Motion,.37 

Chapter V. Balancing of Engines.59 

Chapter VI. The Cylinder,.69 

Chapter VII. Piston, Piston Rings and Piston Pin, 108 

Chapter VIII. Offset Cylinders,.155 

Chapter IX. The Crankshaft, .'.167 

Chapter X. The Connecting Rod,.197 

Chapter XI. Valves and Valve Gears,.218 

Chapter XII. Camshaft and Accessories Drives, . 260 

Chapter XIII. Crank Case and Oiling System, . . 271 

Chapter XIV. Starting Crank, Manifolds and Muf¬ 
fler, .316 

Chapter XV. The Flywheel,.332 

Chapter XVI. Speed Control—The Governor, . . . 352 

Chapter XVII. Power Output and Other Character¬ 
istics, .361 

Chapter XVIII. Sleeve, Piston and Rotary Valve 

Motors,.381 

Chapter XIX. Air Cooling,.397 

Chapter XX. The 'Two Stroke Cycle Motor, . . . 407 

Chapter XXI. Motor Tests,.. . . . 433 

Appendix, ."469 


VII 

















LIST OF PLATES. 


Plate I —Pierce-Arrow Four Cylinder Motor. 

Plate II —Fiat Four Cylinder Motor. 

Plate III —Marmon Four Cylinder Motor. 

Plate IV —Benz Four Cylinder Motor. 

Plate V—Reo Four Cylinder Motor. 

Plate VI—F. N. Four Cylinder Motor. 

Plate VII —Velie 'Four Cylinder Motor. 

Plate VIII —Sunbeam Four Cylinder Motor. 

Plate IX —Continental Six Cylinder Motor. 

Plate X—Carter Four Cylinder Piston Valve Motor. 

Plate XI —Chalmers Six Cylinder Motor. 

Plate XII —Argyll Four Cylinder Single Sleeve Valve Motor. 

Plate XIII —Moline-Knight Four Cylinder Sleeve Valve 
Motor. 

# 

Plate XIV —Weideley Six Cylinder Motor. 

Plate XV —Lozier Four Cylinder Motor. 

Plate XVI —Franklin Six Cylinder Air Cooled Motor. 

Plate XVII —'Northway Six Cylinder Motor. 

Plate XVIII —Rutenber Four Cylinder Motor. 

Plate XIX —Paige Detroit Four Cylinder Power Plant. 
Plate XX —Stellite Four Cylinder Motor. 

Plate XXI —Maxwell Double Cylinder Opposed Motor. 

Plate XXII —Ferro Eight Cylinder Motor. 

Plate XXIII —Piston Displacement Chart. 


VIII 


CHAPTER I. 


THE WORKING MEDIA—GASOLINE AND AIR. 

The process by which mechanical power is generated in 
gasoline motors from the chemical potential energy of the 
gasoline is substantially as follows: Gasoline is mixed with 
seventeen to eighteen times its weight of atmospheric air. in 
which it is vaporized. This explosive mixture is introduced 
into the working cylinders of the motor, where it is ignited 
and burned to carbonic acid gas and water vapor. The gases 
thus formed, together with the nitrogen in the air (which 
does not take part in the combustion), are heated to a high 
temperature by the heat of combustion of the fuel. In this 
way their pressure is greatly increased and they act upon the 
motor pistons, imparting to the pistons a rectilinear motion, 
which by means of the crank shaft and connecting rods is 
transformed into rotary motion. Gasoline and air are there¬ 
fore the working media in a gasoline motor, and before taking 
up the study of motor mechanism it will be well to acquaint 
ourselves with the physical properties of these media. 

Gasoline is a colorless, mobile and relatively light liquid 
which is produced from crude petroleum by fractional distilla¬ 
tion. According to works on the subject, gasoline is that por¬ 
tion of the crude petroleum which distills at 140° to 158° Fahr. 
and which has a specific gravity of 0.636 to 0.70. The product 
distilling at less than 140° Fahr. is known as chymogene, and 
that distilling above 158° as benzine or naphtha. The latter 
distillates include all products passing over between 158° and 
338° Fahr. At the latter temperature kerosene begins to 
distil over. In recent years, as the demand for gasoline has in¬ 
creased, the refiners have marketed heavier and heavier products 
and the greater part of the automobile fuel now used ranges 
above the higher density limit of the above definition for gasoline. 

Gasoline is a mixture of various hydrocarbon compounds, all 
belonging to the same series, known as the paraffin series. The 
principal ones of these compounds are: 


1 



2 


THE WORKING MEDIA. 


Pentane, C 5 Hi 2 
Hexane, C 6 H U 
Heptane, C 7 H )6 

There are several others, both lighter and heavier than those 
mentioned, the general chemical formula being 

C n H 2 n 4 - 2 • 

The chemical diagram of hexane is shown below: 

H H H H H H 

I I I I I I 

H-C-C-C-C-C-C-H 

I I I I I I 

H H H H H H 

and the diagrams for the other members of this group are 
quite similar. 

The most important physical qualities of gasoline are evi¬ 
dently its 

Density or specific gravity. 

Coefficient of heat expansion. 

Latent heat. 

Vapor tension. 

Vapor density. 

Viscosity. 

Calorific or heat value. 

The specific gravity is the ratio of the weight of a certain vol¬ 
ume of gasoline at 6o° Fahr. to the weight of the same volume 
of water at 39 0 Fahr. 

The coefficient of heat expansion is the proportional increase 
in volume per degree Fahr. as compared with the volume at 6o° 
Fahr. 

The latent heat is the amount of heat required to transform one 
pound of liquid gasoline into a dry saturated vapor from and at 
the boiling point. 

The vapor tension is perhaps best explained by stating that if 
gasoline is placed in a closed vessel and heated, the pres¬ 
sure observed while the gasoline is at any particular tem¬ 
perature is the vapor tension corresponding to that temperature. 
This pressure is counted from zero pressure, or is equal to the 
gauge pressure plus 14.7 pounds per square inch. The vapor 
tension is equal to atmospheric pressure, or 14.7 pounds per 
square inch, at the boiling point of the liquid in the atmosphere. 

The vapor density is the ratio of the weight of a certain volume 
of dry gasoline vapor to the weight of an equal volume of dry 
atmospheric air at the same pressure and temperature. 

The viscosity is that quality which determines the rapidity of 





THE WORKING MEDIA. 


3 


flow of the gasoline through a given size of tubular outlet under 
a certain head of pressure. The amount of fluid passing through 
a small tube of a length relatively large in proportion to its bore 
is inversely proportional to the viscosity. 

The calorific or heat value is the amount of heat liberated upon 
burning the gasoline in an excess of air or oxygen. It is ex¬ 
pressed in British thermal units. One British thermal unit is the 
amount of heat required to raise one pound of water from 39 0 to 
40° Fahr. The British thermal unit is equal to 772 foot pounds. 
(778.1 according to the most recent determinations.) 

It will be understood from the start that since gasoline is 
not a definite chemical compound, but a mixture of several such 
compounds in varying proportions, the various properties 
cannot be given with absolute accuracy. The properties of 
the constituent compounds are well known, however, and 
some of them are given in the following table: 

Per Cent, of Per Cent, of Boiling Specific Vapor 


Name. Carbon. Hydrogen. Point—Fahr. Gravity. Density. 

Pentane . 83.20 16.71 87 .640 2.538 

Hexane . 83.68 16.32 143 .676 3.053 

Heptane . 83.90 16.04 165 .718 3-547 


The different grades of gasoline are usually distinguished 
by their density. It has already been stated that the specific 
gravity of gasoline (the ratio of the weight of a certain vol¬ 
ume of it to the weight of the same volume of water at 39 0 
Fahr.) varies between 0.636 and 0.700. It is customary, how¬ 
ever, to express the density in terms of the Baume scale, and 
in order to enable the reader to readily convert Baume de¬ 
grees into specific gravities, and vice versa, a conversion 
table covering that portion of the Baume scale which can pos¬ 
sibly be of use in this connection is given below: 


BAUME-SPECIFIC GRAVITY CONVERSION TABLE. 


Baume Degrees. 

Specific Gravity. 

56 . 


57 . 


58 . 


59 . 


60. 


61. 


62. 


63. 


64. 

.724 

£5. 


66. 

. 7 i 7 

67. 


68. 



Baume Degrees. 

Specific Gravity. 

69. 


70 . 


7i. 


72 . 


73. 


74. 


75. 

.685 

76. 


77. 


78 . 


79. 


80. 

.669 






























4 


THE WORKING MEDIA. 


The Baume and specific gravity scales are connected by the 
following two equations: 

Deg. B = -—-130 at 63.5° Fahr. 

Spec. Gr. J J J 

Spec. Gr. =--at 63.5^ Fahr. 

130 + Degs. B 

Since the standard temperature for making density tests of 
gasoline is 6o° Fahr., instead of 63.5°, 0.002 is added to the 
specific gravity corresponding to any degree Baume as calculated 
by means of these equations. 

The density of gasoline is determined by means of an 
hydrometer, and the proper temperature at which to make 
the test is 6o° Fahr. The reason for this is that gasoline, 
like all other substances, expands when heated and contracts 
when cooled. If, therefore, it is tested at a temperature 
either much greater or much less than 6o°, the reading ob¬ 
tained will be incorrect and a correction factor must be 
applied. The specific gravity will naturally be less at higher 
temperatures and greater at lower temperatures. If 

d\ = density at t° F. 

dgo = density at 6o° F. 

dt = <^6.) L^ 1 — a — 6°)] , 

a being the coefficient of expansion. According to tests 
made by Dr. W. Watson, 
a — 0.00086 for pentane, 
a = 0.00074 I° r hexane, 
a — 0.00055 I° r heptane, 

a = 0.0007 for gasoline (average of nine samples of different 
density.) 

Consequently we may write for ordinary gasoline 
dt = d 60 1 — o. 0007 (t — 60) J 

In ordinary commercial practice it is convenient to express 
the specific gravity of gasoline to two decimal points, and 
since it usually ranges around 0.72, and the coefficient of ex¬ 
pansion 0.0007 is equal to T? Vo » ^ takes a variation of tem- 
1 

perature of =20° Fahr. (approximately) to effect a 

1430 

change of “one point” in the density or specific gravity of 
the gasoline. 

Gasoline vaporizes readily at ordinary atmospheric pres¬ 
sures and temperatures, and it draws its heat of vaporization 
from the surrounding air. Little is therefore said, as a rule, 
regarding its latent heat or heat of vaporization. It is known, 







THE WORKING MEDIA. 


5 


however, that for an average gasoline of 0.70 specific gravity 
this amounts to 140 British thermal units per pound, as com¬ 
pared with 965.7 British thermal units for water. The specific 
heat of the liquid gasoline is 0.45. 

Gasoline of 76° Baume has a vapor tension of 2.53 pounds 
per square inch at 50° Fahr., and for other temperatures this 
tension may be found by the following adaptation of Ran- 
kine’s formula: 


1770 _ 377 - 00 ° t 


log t — 5.264 — 


y-. rp 2 


where t is the vapor tension in pounds per square inch and T the 
absolute temperature Fahr. The vapor tension at a given tem¬ 
perature is, of course, different for different grades of gasoline. 

Gasoline vapor has a high density, and when gasoline is 
spilled on a floor the vapor from it will remain near the floor 
and spread rapidly to a considerable distance. The average 
vapor density is 3.25; that is, at a given temperature and pres¬ 
sure, a unit volume of gasoline vapor weighs 3.25 times as 
much as a unit volume of air. The density of hexane vapor 
is three times that of air. The weight of unit volume of atmos¬ 
pheric air at given temperatures will be given later on, and 
from it the weight of gasoline vapor at different temperatures 
can be calculated. 

Tests to determine the viscosity of various grades of gasoline 
and the change of the viscosity with the temperature have been 
made by J. S. G. Thomas and Dr. W. Watson, and their results are 
given in an appendix to a paper by Dr. Watson on “Thermal and 
Combustion Efficiency of a Four Cylinder Petrol Motor,” which 
is published by the Incorporated Institution of Automobile Engi¬ 
neers (London). The following data are taken from that paper: 
The viscosity is expressed in C. G. S. units, but the figures are 
directly applicable to apparatus constructed according to English 
measures by means of the following formula: The weight of 
gasoline of density A which will flow in t seconds through a fine 
tube of diameter d inch and length l inch (/ being great compared 
to r) when driven by a head H inches, is 



pounds, 


where A is the density or specific gravity of the gasoline; H the 
head in inches; d the diameter of the tube in fractions of an 
inch; t the time in seconds; l the length of the tube in inches and 
v the viscosity. 





6 


THE WORKING MEDIA. 

VISCOSITY OF GASOLINE. 


Density 
at 15 0 C. 

5° C. 

-Viscosity at- 

15 0 c. 

25 ° c. 

0.680*. 


0.00342 

0.00319 

0.684 . 


0.00352 

0.00332 

0.704 . 


0.00380 

0.00359 

0.707 . 


0.00420 

0.00398 

0.714 . 


0.00404 

0.00385 

0.719 . 


0.00420 

0.00398 

0.720 . 


0.00421 

0.00400 

0.721 . 


0.00421 

0.00400 

0.725 . 


0.00454 

0.00430 


* Hexane. 


It will be observed that with a few exceptions the viscosity in¬ 
creases uniformly with the density, and it in every case decreases 
uniformly as the temperature increases. 

The calorific value of gasoline varies slightly with the com¬ 
position, but for practical purposes sufficient accuracy will be 
insured by taking it as 19,000 B. T. U. per pound. The lower 
heating values, as determined by a Junker or similar calorimeter, 
are somewhat lower, and those of the three simple compounds 
found in gasoline are as follows (according to Dr. Watson) : 
Pentane, 18,410 B. T. U. per pound; hexane, 18,770; heptane, 
18,720. It will then be seen that the differences are very slight, 
but the differences are appreciable when the heat units per gal¬ 
lon are considered, owing to the difference in specific gravity. 

Atmospheric Air —Dry atmospheric air is composed of 77 
parts by weight of nitrogen and 23 parts by weight of oxygen. 
The air is capable of holding in suspension considerable 
amounts of water vapor, which have the effect of a diluent. 
In reality the atmosphere contains small quantities of 
other elements, but they are of no consequence in the opera¬ 
tion of a gasoline motor and may be neglected. The oxygen 
of the air combines with the elements of the fuel to form new 
compounds, while the nitrogen does not take part in the com¬ 
bustion. It may be stated that at ordinary atmospheric pressure 
(29.92 inches of mercury) air at a temperature of 32 degrees 
Fahr. weighs 0.0807 pound per cubic foot. Air expands 1-491 of 
its volume at 32 degrees Fahr. for every degree Fahr. increase 
in temperature. 

We are now in position to make a calculation of the amount 
of air required to burn a certain quantity of gasoline. For the 
sake of simplicity we will assume that this gasoline consists 
solely of hexane (GHh). Since the proportion of hydro¬ 
gen and carbon in other hydrocarbon compounds is very nearly 


















THE WORKING MEDIA. 


7 


the same, and since hexane is the chief constituent of all gaso¬ 
lines of a density near 0.70, this assumption is justified. Carbon 
has an atomic weight of 12 and hydrogen of 1, hence the pro¬ 
portions by weight of the two elements in gasoline (hexane) are 

6x12 = 72 parts of carbon to 
14 x 1 = 14 parts of hydrogen, 
or expressing the proportions on a percentage basis, 

83.72 per cent, of carbon 
16.28 per cent, of hydrogen. 

When there is plenty of air mixed with the gasoline, on igni¬ 
tion the carbon of the fuel will combine with the oxygen of 
the air to form carbon dioxide gas (CO2) and the hydrogen 
will combine with the oxygen of the air to form water vapor 
(H 2 0 ). Remembering that oxygen ha? an atomic weight of 
16, we may write the reactions as follows: 

12 C + 32 O = 44 C 0 2 
2 H —j- 16O — i8H z O 

The coefficients in each of these equations show the weight 
proportions in which the elements combine. Now let us take one 
pound of gasoline. We then have 0.837 pound carbon and 0.163 
pound hydrogen, and by using the proportions given in the above 
equations we can find the quantities of oxygen required for the 
combustion of this carbon and hydrogen. 

0.837 C + (ff) 0.837 O = (ff) 0.837 co 2 
o. 163 H + (+) 0.163 O = (V) °-163 H 2 0 

Carrying out the multiplications we obtain 

0.837 C + 2.216 O = 3.053 co 2 
o. 163H-j— i ■ 302 O = i.465H a O 

Total Oxygen required 3.518 

But we saw above that this oxygen is mixed in the atmos¬ 
phere with nitrogen in the proportion of 23:77; consequently, 
3.518 pounds oxygen is mixed with ff X 3.518= 11.774 pounds 
nitrogen or is contained in X 3.518=15.3 (approx.) 
pounds of air. Therefore one pound of gasoline (hex¬ 
ane) theoretically requires 15.3 pounds of atmospheric air 
for its complete combustion. The process of combustion 
in the gasoline engine may therefore be written in chemical 
terms as follows: 

1 pound gasoline -j- 15.3 pounds air 

s - A -\ f --\ 

0.837 C + 0.163 H 3.518 O + 11.774 N 
= 3.053 C 0 2 + 1.465 H 2 0 + 11.774 N. 

Since the mixture of gasoline vapor and air is never perfectly 
homogeneous it is always advisable to have a slight excess of 







8 


THE WORKING MEDIA. 


air in the mixture, from 18 to 20 times the weight of the gaso¬ 
line. If insufficient air is present the hydrogen will burn first, 
and then a portion of the carbon will burn to carbon dioxide 
(CO2) and the rest to carbon monoxide (CO). The com¬ 
bustion to carbon monoxide not only gives off a much smaller 
amount of heat, but carbon monoxide is a very poisonous gas 
and makes a foul exhaust. 

The higher heat value is the amount of heat obtained by burn¬ 
ing one pound of gasoline in air and cooling the resulting products 
down to 6o° Fahr. The water vapor in the products of combus¬ 
tion is then condensed and the latent heat or heat of condensa¬ 
tion of the water vapor is thus added to the result. The lower 
heat value is the result obtained by cooling the products of com¬ 
bustion to 212 0 without condensing the vapor. The theoretical dif¬ 
ference between the two heat values can readily be calculated. 
We have found that in burning one pound of gasoline 1.465 
pounds of water vapor is formed. The latent heat of water is 
965.7 B. T. U. per pound, and the cooling of the water from 212 0 
to 6o° gives a further 152 B. T. U. per pound, making a total of 
1,118 B. T. U. Of the other two constituents of the products of 
combustion, carbon dioxide and nitrogen, the former has a specific 
heat of 0.216 and the latter of 0.244. The total heat gained in 
condensing the water vapor and cooling all of the products of 
combustion of one pound of gasoline from 212 0 to 6o° Fahr. is 
therefore 


h 2 o 

1.46 X 1118 = 

1632 

co 2 

3.05 X 152 X 0.216 = 

IOO 

N. 

11.77 X 152 X 0.244 = 

436 


Total.2,168 B. T. U. 

This is the difference between the lower and higher calorific 
values per pound. 

It is evidently the most logical plan, in making calculations of 
thermal efficiency, to figure with the lower heat value of the fuel, 
as it is physically impossible to utilize the latent heat of water 
vapor in a gasoline engine. It is the custom in this country to 
base calculations on the lower heat value. 

The total heat liberated in the combustion of gasoline is the 
sum of the heats liberated by the combustion of the carbon and 
the combustion of the hydrogen, minus the heat required to 
break up the hydrocarbon molecules. In the combustion of 
carbon to carbon dioxide the heat liberated per pound of the 
carbon dioxide is 3,860 B. T. U. In t/he combustion of hy¬ 
drogen to water vapor the heat liberated per pound of vapor 
is 5.830 B. T. U. The dissociation of the gasoline or hexane 






THE WORKING MEDIA. 


9 




into hydrogen and carbon requires 420 B. T. U. per pound. If, 
now, we take the quantities of carbon dioxide and water vapor 
produced by the combustion of one pound of hexane as found 
above, we may write the following heat balance: 

3.053 X 3,860 = 11.784 
1.465X5,830= 8,541 

20,325 
— 420 


19,905 

This result, 19,905 B. T. U. per pound of gasoline, is slightly 
higher than the best experimental results. 

Gasoline mixtures containing the relative proportions of 
constituents required for complete combustion are readily 
ignited by an electric spark. When the proportions of gaso¬ 
line and air vary much from this standard either way it be¬ 
comes more difficult to ignite the mixture. Greater heat is 
required to effect the ignition, and once ignited the charge 
burns more slowly. The limiting proportions for which igni¬ 
tion is still possible at atmospheric pressure are about 1:7 
(rich mixture) and 1:32 (lean mixture). 


\ 




CHAPTER II. 


THE LAWS OF GASES—THERMODYNAMICS. 

Heat is a form of energy and is convertible into other forms 
of energy, which in turn are convertible into it. Energy is de¬ 
fined as the capacity for performing mechanical work. Experi¬ 
ments conducted by Helmholtz, Joule and Mayer during the past 
century proved that energy is indestructible; it can be converted 
from one form into another and reconverted any number of 
times, but the total amount of it never changes. The most 
familiar form of energy is mechanical energy, which is measured 
in foot-pounds. One foot-pound is the energy expended in rais¬ 
ing one pound of matter one foot against the force of gravity. 
Heat energy is measured in British thermal units. The B. T. U., 
as already stated, is that amount of heat which is required to 
raise one pound of water from 39 to 40 degrees Fahrenheit. 
Joule’s experiments showed that 

1 B. T. U. = 772 foot-pounds, 

and the figure 772 is known as Joule’s equivalent or the mechani¬ 
cal equivalent of heat. It is represented in mathematical discus¬ 
sions by the letter J. 

There are two broad principles underlying the conversion of 
heat into mechanical energy and the reciprocal process, known 
respectively as the first and the second law of thermodynamics. 
They may be stated as follows: 

First Law of Thermodynamics: When work is done by 
the expenditure of heat, the quantity of heat consumed is a 
measure of the quantity of work done, or of the energy acquired 
in a new form; and conversely, if work is converted into heat 
energy the conversion takes place in a definite ratio. 

Second Law of Thermodynamics: If the total actual heat 
of a homogeneous and uniformly hot substance be conceived to 
be divided into any number of equal parts, the effects of these 
parts in causing work to be performed are equal (Rankine). 


10 



THE LAWS OF GASES. 


11 


Perfect Gases —It has been found desirable in scientific dis¬ 
cussions to distinguish between perfect gases—those far from 
their temperature of liquefaction—and imperfect gases or vapors 
—those near their point of liquefaction. Although gasoline vapoi 
forms one of the constituents of the combustible charge used 
in gasoline motors, it is convenient in discussing the theory of 
these motors to consider the charge as a perfect gas. 

When a quantity of a perfect gas is confined in a closed vessel 
it exerts a uniform pressure over the entire interior surface of 
the vessel. This pressure is known as the gas pressure and is 
usually expressed in pounds per square inch. If the volume 
occupied by the gas is changed, the pressure will generally be 
changed. It has been found that if the volume of the gas is re¬ 
duced in a certain proportion—the temperature of the gas remain¬ 
ing the same—the pressure will be increased in the same propor- 
. tion; inversely, if the volume is increased in any proportion, the 
temperature of the gas remaining the same, the pressure will be 
reduced in the same proportion. This law is known as Boyle’s 
or Marriotte’s law, and is mathematically expressed by the 
equation 

P V = Constant, 

where P is the pressure and V the volume. Compression of the 
gas not accompanied by any change in the temperature of the 
same is known as isothermal compression, and expansion under 
the same conditions as isothermal expansion. 

Ordinarily, however, the compression and expansion of gases 
are accompanied by temperature variations, as will be explained 
farther on. 

Absolute Temperature Scale —When a gas is heated, if it 
is free to expand, it will do so, and a certain relation has been 
found to exist between the volume increase and temperature 
increase. It is known that the ordinary thermometric scale is en¬ 
tirely arbitrary, the freezing point of water being called 32 de¬ 
grees, the boiling point of water 212 degrees and the entire scale 
divided into parts equal to 1/180 of the distance between these 
two points. But if we fix our zero at what on v the ordinary 
Fahrenheit scale would be —461° and measure temperatures 
from that point, then the increase in volume of a perfect gas 
(the pressure remaining constant) will be directly proportional 
to the increase in the temperature of the gas, or the increase in 
pressure (the volume remaining constant) will be directly pro¬ 
portional to the increase in temperature. 

This point on the temperature scale, viz., —461° Fahr., is 
.ailed the absolute zero, and temperatures measured from it are 


12 


THE LAWS OF GASES. 


called absolute temperatures. The above relationship in its most 
general form is known as Charles’ law and is expressed mathe¬ 
matically as follows: 


P V' 
T ‘ 


Constant , 


P being the absolute pressure; V, the volume, and T the abso¬ 
lute temperature. 

Specific Heats—The amount of heat required to raise the 
temperature of any substance by one degree Fahrenheit depends 
upon what is known as its specific heat. It has already been 
pointed out that it requires one British thermal unit to raise one 
pound of water from 39 to 40 degrees Fahrenheit, and water is 
considered to have _a specific heat of unity; hence the specific 
heat of other substances is equal to the amount of heat expressed 
in British thermal units required to raise the temperature of one 
pound of these substances one degree Fahrenheit. In the case 
of gases, however, it makes a difference whether the rise in tem¬ 
perature is accompanied by a rise in the pressure of the gas 
or by an increase in the volume. If the gas is heated at constant 
volume, the heat in British thermal units required to raise one 
pound 1 degree Fahrenheit is known as the specific heat at con¬ 
stant volume (C v '), while if the gas is heated at constant pres¬ 
sure the heat in thermal units required to increase the tempera¬ 
ture 1 degree Fahrenheit is known as the specific heat at constant 
pressure (C ). The specific heat of air at constant volume has 
been found by experiment (Regnault) to be 0.169. From this we 
can calculate the specific heat of air at constant pressure as 
follows • 

Consider a cylinder of such a bore that its cross section meas¬ 
ures just one square foot. A piston rests in this cylinder with 
its head exactly one foot from the cylinder head. The space 
back of the piston measures then exactly one cubic foot, and we 
will consider this space filled with air at atmospheric pressure 
and at a temperature of 62° Fahr., or 523 0 absolute. We will as¬ 
sume that the piston has an air-tight fit in the cylinder, but can 
move in it without friction. 

Now let heat be applied to the air confined in the cylinder, 
and its temperature thereby be raised. As the temperature rises, 
the pressure of the air tends to rise, but we have assumed that 
the piston moves in the cylinder without friction; and, since its 
outer face is subjected only to atmospheric pressure, immediately 
the pressure of the confined air rises above atmospheric pressure 
the piston moves outward, thus keeping the pressure within down 



THE LAWS OF GASES. 


13 


to atmospheric. If we keep on supplying heat to the confined 
gas until its temperature is 585° Fahr., or 1046° absolute, accord¬ 
ing to the law of Charles the volume will be doubled.. The 
piston will then have moved exactly 1 foot, making the volume 
filled by the air back of it 2 cubic feet. The atmosphere is all 
the time pressing against the outer face of the piston with a 
pressure of 14.7 pounds per square inch, or a total pressure of 

144X14.7 = 2,116.8 pounds, 

and since the piston was moved 1 foot against this pressure the 
energy expended in thus “pushing back the atmosphere” is 
2,116.8 foot-pounds. This is equal to 

2,Il6.8 r> 'T- ti 

—-= 2.742 B. T. U. 

772 

Now, one cubic foot of air under atmospheric pressure at 62° 
Fahr. weighs 0.0761 pound. Consequently, in order to raise 
0.0761 pound of air 523 0 Fahr. at constant pressure requires an 
expenditure of energy for pushing back the atmosphere of 2.742 
B. T. U. The amount of energy required per pound of air per 
degree Fahrenheit is therefore 

-^ 24 ?-. = 0.0688 B. T. U. 

0.0761 X 523 

Adding this to the energy required to raise the temperature of 
one pound of air one degree Fahrenheit at constant volume, we 
find for the heat required to raise it one degree at constant 
pressure 

0.1691 -f- 0.0688 = 0.2379 T. U. 

We have therefore: 

Specific heat of air at constant volume CV = 0.1691 ; 

Specific heat of air at constant pressure C P , = 0.2379. 

The ratio of these two factors 

2^379 = 1.408 

o. 1691 

plays quite an important part in thermodynamics. 

Consider again the aforesaid cylinder with a cross section of 
1 square foot, and containing a volume V of air at the tempera¬ 
ture, T, and pressure, P (both absolute). Now, let heat be ap¬ 
plied to this gas until the temperature becomes Ti and the volume 
Vi. Part of the heat is expended in raising the temperature of 
the air and the rest in pushing back the atmosphere or other 
pressure holding the piston in place; that is, in doing external 
work. We will consider the quantity of air in the cylinder a unit 
quantity and P the pressure per square foot. As regards units, 
we can, of course, make them anything we choose, only we must 





14 


THE LAWS OF GASES. 


use the same units throughout the discussion. The foregoing are 
here chosen in order to eliminate coefficients which would simply 
complicate the work, but would not affect the result. 

The heat required to raise the temperature of the gas in this 
operation is the same as would be required to raise the tem¬ 
perature the same amount at constant volume, viz., 

C v (Tt - T ). 

The total heat required being C p ( T\ — T), it follows that the 
difference (C p — C v ) (Ti — T ) is expended in doing external 
work. This work, however, can also readily be expressed in 
otb^r terms. Since the cross section of the cylinder is exactly 
I square foot, the volume increase in cubic feet is the same as 
the motion of the piston in feet, and vice versa. Therefore, 
when the volume increases from V to Vi the piston moves 
V\ — V feet, and since it moves against the pressure P, the 
work done is 

P ( V\ — V) foot-pounds. 
which is equivalent to 


J 


(V 1 — V) heat-units , 


J being the mechanical equivalent of heat. 

Equating the two expressions for the external work done we have 

P 


(C P — Cy ) ( T x — T) = ~ (Vi — V) . 

But since the pressure P remains constant, we have 

V Vj 


(i> 


or 


- 1 (Charles’ law), 

T T x 


V 

V l = T 


Substituting this value of V x in equation (i) we obtain 

(Cp — C v )(r,— r) = ^ (t,£- vy 

Dividing both sides by 

(SH: 

we have 

’ (Cp-C. )z = ^r. 

which by transformation gives 

PV =J(C f — C,). 


But by Charles’ law 


T 


P V 

— Constant t 




THE LAWS OF GASES. 


15- 


consequently, 

J {C p — Cv ) — Cotisiant. 

This constant is generally designated by R, so that we may- 
write : 


P V 

y ~ J( ^ p Cv) = P .(2): 

It will be noted that in dealing with gases there are three 
main factors, viz., the volume, the temperature and the pressure. 
We have thus far always assumed one of these factors to be 
constant and then investigated the change of the second with a 
change in the third. But in practical work with gases all of the- 
foregoing factors often change simultaneously. 

Adiabatic Changes of State—According to the principle of 

the first law of thermodynamics, if heat is supplied to a quantity 
of air or gas the amount of heat thus supplied will be equivalent 
to the heat used for raising the temperature of the air 01 gas 
plus the energy expended in doing external work. Considering., 
therefore, a small amount of heat, H, supplied, which will cause 
an infinitesimal increase, d T, in the temperature of the air andl 
an infinitesimal increase, d V, in the volume, we may write: 

//= Cv 

J 

From this it readily follows that if any changes take place in- 
the state of the air without the total amount of heat in it being 
either added to or subtracted from (H = o), then 


CvdT + 


PdV 

J 


o. 


Such changes in the state of gas, not accompanied by the addi¬ 
tion of heat from outside, or the subtraction of heat from the 
gas, are known as adiabatic (equal heat) changes. 

The limiting ratio of temperature to volume change is therefore 


d T _ _ P 
dV~ JL\ 

We found that 

— = R=J(C — Cv), 

T 

from which it follows that 

PV— TR= TJ(C p —Cv). 
Differentiating with respect to P, V and T, we have : 

PdV-\- V d P — J(C P — Cv)dT * 

j± T (C p —Cv) = P+ V~- 
J dV db 


( 3 )J 











16 


THE LAWS OF GASES. 


d 'T 

Substituting the value offound in (3), we have 

a l 

— p( -f- P = P+ F —> 

\Cr dV 

which can be changed to 

dP / C P \ dr, 

P \Cy ) V 

Now, it may be found in text books on the Integral Calculus 
that the integral of a differential expression of the form — is 

X 

log x. Hence, integrating, we have: 


transposing, 


log P — 


— log V -f- Constant. 
Cv 


log P + 



log V= Constant , 


and taking antilogs, 


Cp 

p = Constant 


The ratio of the specific heats 


Cp 

Cv 


we may write : 


is generally designated by 7, so 


P V^ = Constant ...(4) 

This gives us an expression connecting the pressure and volume 
for adiabatic changes in the state of the gas. The temperature 
is not a factor in this equation, but an expression for it may 
readily be found as follows: 

From the equation, 


we find that 


P V 
T 


— P, 


P- 


TR 

- 5 

V 


and multiplying both sides by 

P PRjd = constant. 


or, since R is a constant 

T v'Y = Constant . (5) 

We also may write the foregoing equation in the form 









THE LAWS OF GASES. 


17 


so that 



and multiplying both sides by P , 


PV 1 =zP 

from which it follows that 



y 

= Constant 


T y 

“ ' = Constant .(6) 

pi — 1 v ' 

Summarizing, for adiabatic expansion or compression of gas 
the following three equations hold: 


P = Constant; 

T V^ — Constant; 

T 1 

-= Constant . 

py-1 

The following table gives the values of C P , Cv and 7 for the 
different gases involved in gasoline motor work: 


Gas. . C p C v 7 

Air . 0.2375 0.1684 1.408 

Oxygen . 0.2175 0.1551 1.370 

Nitrogen . 0.2438 0.1727 1.412 

Hydrogen . 3*409 2.4110 1*414 

Carbon monoxide. 0.2450 0.1736 1.411 

Carbon dioxide. 0.2869 0.172 1.668 

Steam . 0.4805 0.370 1.298 


Temperature of Gasoline Combustion—We found that 
one pound of gasoline in burning evolves approximately 19,000 
B. T. U. This heat energy in a gasoline motor as now com¬ 
monly used is applied to the gas at constant volume, and all of 
it therefore goes to raise the temperature of the gas. The three 
constituents of the gases in the cylinder after the charge has 
been burned are nitrogen, carbon dioxide and water vapor, and 
considering the proportions in which they are present and the 
specific heat at constant volume of each, we should expect the 
specific heat of the products as a whole to be about 0.20. Assum¬ 
ing that one part of gasoline vapor is mixed with 20 parts of 
air, we should expect a rise in temperature on combustion of 

— I 9 ’ OOQ — = 4,500° Fahr. {a$pr.). 

21 X 0.20 


In reality the temperature never exceeds 3,000 degrees Fahren¬ 
heit. Various reasons have been advanced in explanation of this 
discrepancy between theory and practice. The first and most 
plausible explanation is that the specific heat of gases, instead 













18 


THE LAWS OF GASES 


of being constant, increases with the temperature. This subject 
has been experimentally investigated by Mallard and Le Chate- 
lier. From the results of their experiments it is found that there 
is a linear relation between the specific heat at constant volume 
and the temperature of the gas, and that the specific heats may 
be represented by the following equations: 

For C 0 2 , Cm = 0.1408 -f- 0.0000463 T; 

For HoO, C y = 0.279 -j- o.oooioii T; 

For N 2 , Cm = o. 171 -f- 0.0000120 T; 

For 0 2 , Cm — 0.150 -J- 0.0000104 T. 

In these T is the regular Fahrenheit temperature and C v the 
average specific heat up to the temperature T. It will be 
noted that the temperature coefficient for those gases which 
have the lowest temperature of liquefaction (N 2 and 0 2 ) are 
the lowest; in other w r ords, for these gases the specific heat is 
most nearly constant. If we determine the temperature of 
combustion by making use of the foregoing specific heats, the 
result is still about 3,700 degrees Fahrenheit, and the increase 
in the specific heats at high temperature, therefore, does not 
fully explain the discrepancy between calculated and observed 
results. 

Other explanations are that the cooling effect of the walls 
prevents the gas from attaining the calculated temperature, that 
compounds formed by the combustion, viz., carbonic acid gas 
and water vapor, become dissociated at high temperatures, 
whereby heat energy is absorbed, and that the chemical re¬ 
actions are not yet terminated at the moment the maximum 
temperature is reached, so that this temperature does not depend 
upon the total heat of combustion of all of the gas. This 
latter theory is referred to as that of “after burning.” The 
fact that the pressure and temperature of explosion are higher 
in a motor with hemispherical combustion chamber than in a 
motor with valve pockets on opposite sides proves that the cool¬ 
ing influence of the walls and “after burning” affect the tempera¬ 
ture of explosion. 


CHAPTER III. 


THE OTTO CYCLE. 

The great majority of automobile motors are operated on what 
is known as the Otto or four-stroke cycle. By a cycle is meant 
the succession of operations in a cylinder with a single charge of 
explosive mixture. The Otto or four-stroke cycle comprises the 
following four operations, succeeding one another in the order 
given: 

Admission of the charge to the cylinder. 

Compression of the charge. 

Combustion of the charge (which includes its ignition and ex¬ 
pansion). 

Exhaustion of the products of combustion. 

Each of these four operations occupies the time ot one piston 
stroke. This cycle was originally proposed by Beau de Rochas, 
a Frenchman, who in 1862 took out a French patent on an engine 
employing it. Its first practical application, however, is due to 
Dr. N. A. Otto, a German engineer, who in 1876 patented, and 
in 1878 exhibited at the World’s Fair in Paris the first four- 
stroke cycle engine of the general type in common use today. 

Admission —The first stroke of the Otto cycle is known as the 
admission or intake stroke. The piston then moves outward in 
the cylinder, thereby creating a rarefaction of the air or gas in 
the outer or working end of the cylinder (generally known as 
the combustion chamber). At some point in the wall of the 
combustion chamber is located a valve which, at the proper 
time, places the combustion chamber in communication with the 
source of combustible mixture, the carburetor. During the 
admission stroke there is, as we have seen, a rarefaction of 
the gas in the combustion chamber; in other words, a partial 
vacuum, so that the pressure on the admission valve from the 
inside at this time is less than atmospheric pressure; and since 
the atmosphere presses against the valve from the outside it is 
possible to open the valve by means of this difference in. pres- 


19 



20 


THE OTTO CYCLE. 



THE HORSELESS ABE' 


Fig. i—Admission. 


sure on its two sides, that 
is, by the suction effect. 

At the end of the inlet 
stroke the suction ceases, 
and a light spring will 
suffice then to close the 
valve. This type of inlet 
valve, known as the auto¬ 
matic inlet valve, was very 
extensively used at one 
time, but it has now been 
discarded for automobile 
work, because it is noisy in 
operation and because it 
closes too early when the 
motor runs slowly, and too 
late when the motor runs very fast, thus making it impossible 
to get full charges into the cylinder under these conditions. By 
a full charge is meant a volume of charge equal to the piston 
displacement volume at atmospheric pressure and temperature. 
At present a type of valve is generally used which is positively 
opened by mechanical connection with the motor crankshaft, and 
also positively closed. This valve begins to open a little after 
the beginning of the admission stroke and closes a little after 
the admission stroke has been completed. 

When the motor is rotating at low speed the degree of vacuum 
or suction in the cylinder during the admission stroke will be 
very small, and it will, under any conditions, be the smaller the 
greater the cross section of the inlet ports and passages in pro¬ 
portion to the cross section of the cylinder. The least reduc¬ 
tion in the density of the gas inside the combustion chamber will 
immediately create an inflow of combustible mixture, which will 
all but fill the vacuum. At high engine speeds, however, there is 
considerable resistance to the flow of the gaseous mixture 
through the carburetor, inlet pipe and valve ports, and the cyl¬ 
inder then receives considerably less than a full charge. 

Compression—Shortly after the completion of the inlet stroke, 
the inlet valve is closed, and during the rest of the return stroke 
of the piston the charge is compressed in the combustion 
chamber. When the piston has reached the end of its return 
stroke—the compression stroke—the gas is compressed into from 
one-third to one-fifth or less of the volume it occupied at the 
beginning of the stroke. The space which it then occupies is 
known as the compression space. The object of compressing the 
charge previous to ignition is as follows : 
























THE OTTO CYCLE. 


21 


Upon ignition the whole 
charge is immediately 
brought to a high tem¬ 
perature and a high pres¬ 
sure. By reason of its 
pressure the gas acts on 
the piston and thus con¬ 
verts heat energy into 
mechanical energy. How¬ 
ever, owing to the con¬ 
tact of the highly heated 
gases with the relatively 
cool walls of the combus¬ 
tion chamber, heat is lost 
rapidly. This loss of heat, 
of course, represents a 
loss of energy, as the gases on cooling decrease in pres¬ 
sure and in their capacity for performing mechanical work. 
Other things being equal, this loss of heat is directly propor¬ 
tional to the area of the cooling surface with which a given 
quantity of charge is in contact. By compressing the charge 
before ignition the area of the surface with which it is in con¬ 
tact during ignition and expansion is lessened, and the heat 
losses are reduced. An engine in which the charge is compressed 
previous to ignition is, therefore, more efficient, transforming a 
greater percentage of the heat energy of the fuel into mechanical 
energy. Besides this, the working pressure is greater, and an 
engine of given size will have a much greater output when 
working with compression of the charge than when firing the 
charge at atmospheric pressure. In brief, compression increases 
the fuel efficiency and reduces the weight and bulk of the engine. 

Ignition and Combustion—At or near the end of the com¬ 
pression stroke an electric spark is produced in the compression 
chamber which ignites the charge. If ignition of the whole 
charge took place instantaneously the end of the stroke would 
be the proper time for the spark to occur. It has been shown, 
however, that it takes some time for the flame to be propagated 
through the charge, and for this reason when the motor is run¬ 
ning at relatively high speed the spark must occur slightly before 
the dead centre for the best results. 

Upon ignition the pressure suddenly rises to between four 
and five times what it was previously, and the piston is forced out 
under this pressure. The pressure multiplication factor varies 
with the composition of the charge, the form of the compression 
chamber and other variables. The actual pressure in a gasoline 

























22 


THE OTTO CYCLE. 



Fig. 3.—Expansion. 


motor upon ignition is 
generally between 250 and 
350 pounds per square 
inch, when the motor 
works at full power. The 
pressure drops, however, 
very rapidly, owing both to 
the increase in the volume 
of the gases as the piston 
moves outward and to the 
abstraction of heat from 
the gases by the cylinder 
walls. When the power 
stroke is almost completed 
the exhaust valve begins 
to open and the spent 
gases, which are still under about 50 pounds pressure per square 
inch, escape rapidly. 

Exhaust—The exhaustion of the burnt gases continues all 
through the following return stroke of the piston, the exhaust 
valve remaining open throughout this stroke. During the first 
part of the exhaust period the gases are forced out by their 
proper pressure, which causes them to expand when no longer 
confined, but during the last portion of this stroke the piston 
practically sweeps the gases before it and pushes them out of 
the cylinder. The piston, however, does not sweep through the 
whole of the cylinder space and cannot, therefore, clear the 
compression space of spent gases. This space remains always 
filled with the product of the previous combustion, and is for 
this reason also known as the dead space, or clearance 

space. At the end of the 
exhaust stroke, or slightly 
later, the exhaust valve 
closes, and at about the 
same time the inlet valve 
opens again and the cy¬ 
cle starts anew. In the 
sketch, Fig. 5, is represent¬ 
ed a cylinder of a gasoline 
engine with its piston in 
place. Into the piston head 
is secured a pipe leading 
to a device known as an 
indicator. This consists 
of a cylinder with a piston 
in it. with a coiled spring 



TW. HOR&utaOua 


Fig. 4.—Exhaust. 














































THE OTTO CYCLE. 


23 


on top which tends to force the piston down to the bottom of 
the cylinder. The piston rod extends through the head of the 
cylinder and connects to a set of four links, joined so as 
to form a parallelogram. One member of this linkage is 
extended downwardly and is pivoted at its lower end to a 
fixed bracket. The top member is also extended and carries 
at its outer end a pencil which is lightly pressed against a 
paper card carried on a frame on top of the cylinder. 

The indicator cylinder is always in direct communication 
with the motor cylinder, and the same gaseous pressure that 



Fig. 5. 


acts on the motor piston acts on the indicator piston. The 
indicator piston will thereby be forced upward against the 
pressure of the spring. It is a law of coiled compression springs 
that when subjected to pressure they compress a distance which 
is directly proportional to the pressure. Hence, at any point 
in the stroke of the motor piston the indicator piston will be 
moved up in its cylinder a distance directly proportional to the 
pressure in the motor cylinder. The pencil will be moved a 
vertical distance directly proportional to the motion of the in¬ 
dicator piston, and hence the vertical height of the pencil is 











































24 


THE OTTO CYCLE 


always a measure of the pressure acting in the motor cylinder. 
The pencil is also moved horizontally back and forth, owing to 
the fact that it is carried by the motor piston, and its horizon¬ 
tal position, therefore, always corresponds to the position of 
the motor piston head; in other words, to the point in the 
stroke. 

If the motor piston is moved back and forth in the cylinder 
while the valves are held open, the pressure in the cylinder will 
remain atmospheric, and the pencil will describe a straight hori¬ 
zontal line A A, known as the atmospheric line. If, however, 
the motor performs its regular cycle and the inlet valve closes 
at the end of the inlet stroke, then during the following re¬ 
turn stroke the gas will be compressed and the pencil will de¬ 
scribe curve A B, known as the compression curve. At the 
end of this stroke the charge is ignited and the pressure in the 
cylinder suddenly rises, in consequence of which the pencil de¬ 
scribes a vertical or nearly vertical line B, C. Then, as the piston 
performs the next or power stroke, the pressure in the cylinder 
falls, first rapidly and then gradually more slowly, as indicated 
by the curved line C D, which is known as the expansion 
curve. At the end of the power stroke, as the exhaust valve is 
opened, the pressure drops suddenly to almost atmospheric, as 
indicated by the short vertical line DA. 

These various straight and curved lines form together what is 
known as an engine diagram, which gives an accurate picture 
of what is going on inside the motor cylinder. Such a diagram 
cannot actually be taken from an automobile engine by the ap¬ 
paratus here shown, as the inertia of the moving parts would 
interfere with the results, but the sketch serves very well to ex¬ 
plain what is meant by an indicator diagram. 

Compression Calculation—At the present time it is custom¬ 
ary to work at from 75 to 90 pounds compression absolute. 
The higher the compression the greater the power obtained 
from a motor of given cylinder dimensions, and the higher the 
fuel economy, but when the compression exceeds a certain 
limit there is danger of trouble from overheating of the cylinder, 
self-ignition (ill-timed) due to the heat of the compression, and 
knocking in the bearings. Some years ago automobile motors 
were frequently built with compressions as high as 100 pounds 
per square inch, but lower compressions were found advan¬ 
tageous. 

The compression of the charge approximates an adiabatic 
compression, the expression for which, as we have seen, is 


THE OTTO CYCLE. 


25 


P V = Constant. 

If Pi be the pressure in the cylinder when the compression 
stroke begins; Vi, the volume occupied by the gases at the be¬ 
ginning of the compression stroke (volume swept through by 
piston plus compression space volume) ; P 2 , the pressure at the 
completion of the compression stroke and Vi the volume occu¬ 
pied by the gases at the completion of that stroke (compression 
space volume), then, evidently, 

P x V x y = P 2 V 2 7 = Constant. 

From which it follows that the compression pressure 



The initial pressure Pi varies with the piston speed, the re¬ 
lative size and form of the passages, and with other factors. At 
low engine speeds it is generally very close to atmospheric pres¬ 
sure or 14.7 pounds per square inch, while at normal engine 
speeds this pressure is not more than from 12 to 13 pounds per 

square inch. The ratio -p is known as the compression ratio. 

As regards the value of the exponent 7, this has been de¬ 
termined from actual indicator diagrams and has been found 
to be 1.3 approximately. Assuming, for example, an initial 
pressure of 13 pounds per square inch and a compression ratio 
of 4 to 1, we find the compression pressure to be 
13 X 4 1-3 = 78.8 'pounds absolute. 

Chart I herewith instantly gives the compression pressure for 
any initial pressure from 12 to 15 pounds per square inch, and 
any compression ratio from 3 to 5. 

It is obvious that the compression pressure of an engine of 
given construction is not fixed but varies with the speed, since 
the speed affects the degree of filling, and, hence, the initial 
pressure. In giving the specifications of an engine it is better, 
therefore, to give the compression ratio—the ratio of the volume 
of piston sweep plus the compression volume, to the compression 
volume—as this is fixed. 

Work of Compression —The compression of the charge in 
the compression chamber requires the expenditure of a certain 
amount of work. Let 

b- Cylinder bore in inches; 

/ = Length of stroke; 
r = Ratio of compression. 


26 


THE OTTO CYCLE. 



Chari I. 




























































































27 


THE OTTO CYCLE. 

Then, considering the compression space in the form of an ex¬ 
tension of the cylinder, its height will be 

l 

__ • 

r — i 

The length from the cylinder head to the piston head when the 
latter is at the beginning of the compression stroke is 



Suppose that at the beginning of the compression stroke the 
pressure in the cylinder is P\. The total pressure against the 
piston from the inside of the cylinder then is 

7T b* n . 


4 



Now consider the piston to be at a distance x from the cylin¬ 
der head. The internal pressure then is 

1.3 




x 


_l -3 . 


If the piston is now moved an infinitesimal distance — dx to the 
left against this pressure, the work done will be 


d JV C = 


z^xr, 

4 


(rfh) 


1 3 































THE OTTO CYCLE. 


28 


This is a differential equation of the form 

d y — a x m d x, 

the integral of which is 

m + x 
X 


y — a 


-f- Constant. 


m -(-1 

However, since we integrate in this case between limits, the con¬ 
stant vanishes. The limiting values of x are 


and 


r l 


r — i 


r — i 


Hence 


//(r —i) 


( 8 ) 


Since 


(riy" = Jd-, 

V 1/ — 1/ r — i 

multiplying the part outside the square brackets by 


( f ^.) 


— 0.3 


and dividing the part inside the brackets by the same expres¬ 
sion, we have 

7r b 2 


JV C = 


'jPi_ ( rl \ ["/A - 03 _ 

0.3 \r — 1/ L W j’ 


which may be simplified to 

b 2 P, ( r l 


W c = 


7 T 


, 0.3 


I . 2 


- 1 )' 


This equation gives the result in inch-pounds, if the pres¬ 
sure P\ is given in pounds per square inch and b and l are given 
in inches. In order to get it in the more usual foot-pounds, we 
divide by 12, which gives 

ir b 2 P\ ( rl\ I \ 

I'Ve— “ _ I J ^r°' s — 1 j foot-pounds .(10) 

This, of course, does not take account of the fact that the atmos¬ 
phere presses with a pressure of 14.7 pounds per square inch 
against the outside of the piston, and thus helps to compress the 
gas. The amount of work given by the equation would be re- 

























THE OTTO CYCLE. 


29 


quired to compress the charge if the cylinder were placed in a 
vacuum. It is permissible to neglect the effect of the atmos¬ 
pheric pressure, because the energy thus derived during the 
compression stroke must be expended again during the following 
expansion stroke. However, if it is desired to determine the net 
work of compression, this can easily be done by subtracting from 
the result of equation (io) the energy derived from the atmos¬ 
phere during the compression stroke, which is 


7T 6 2 / 14.7 , 

v o — foot-pounds . 

4 A 12 

Work of Expansion —If after the completion of the compres¬ 
sion stroke the charge is allowed to expand again, the work done 
by the expansion will be the same as that expended in compres¬ 
sing the charge. Equation (10) would therefore apply to the ex¬ 
pansion as well as to the compression if the gas were expanded 
in the same state in which it was compressed. However, at the 
end of the compression stroke the charge is burned, whereby its 
pressure is multiplied several times. Calling this pressure ratio 
a, we may write for the work of expansion: 


Tr^a PJ r l \ I °- 3 \ 

14^ \ r -') foot-founds. 


The difference between the work of expansion and the work 
of compression is evidently the useful work of one cycle (neglect¬ 
ing the slight loss due to suction and exhaust back pressure). 
This useful work of one cycle is: 


IVu= 7r I4 * “ (a — {r — 'ij ( r °' 3 ~ r ) foot-pounds... .(n) 

This equation clearly shows the effect of the compression ratio r 
on the work per cycle, and consequently the variation of the 
output of an engine at constant speed with varying compression. 
For r = 3 the expression 

(rh) {r z ) =0 *5 8 5* 

while for r — 4 

o-s \ 

— ij =0,687, 

which plainly shows the increase in power with increasing com¬ 
pression ratios. 

It is desirable, however, to have an expression for the work of 
the expansion based upon the pressure in the cylinder at the 
beginninar of the expansion stroke (explosion pressure), rather 
than upon the pressure in the cylinder at the beginning of the 










30 


THE OTTO CYCLE. 


compression stroke. Besides this, a study of indicator diagrams 
shows that the exponent 7 is somewhat greater for the expansion 
than for the compression and may preferably be taken as 1.32 
Then, since 


Pi 




1-32 


(Pc being the absolute compression pressure), introducing the 
pressure multiplying factor a and making a few simplifications, 
we may write equation (8) : 


JK _ TTb 2 a Pc ( 1 ) ! ' 32 [( 1 ^~°‘ 32 _ ^ lr 

Dividing the part outside and multiplying the part inside the 
brackets by 

0-32 


we get 


W i 


tt Pa Pc 
14.4 


(^) 

/— ^foot-pounds .(12) 


I 


1 


1 


1 




Fig. 7.—Ideal Diagram. 




















THE OTTO CYCLE. 


31 


Exhaust and Suction Curves—Fig. 7 represents ideal com¬ 
pression and expansion curves. The actual curves differ slightly 
from these curves, for several reasons. In the first place, the 
entire gas in the combustion chamber is not ignited instantly, 
and, therefore, we do not have in an actual diagram a straight 
vertical line Pe P c making very distinct junctions with the com¬ 
pression and expansion curves. If the ignition spark is produced 
while the piston is at the end of the stroke the line Pc Pe will be 
slightly inclined to the right, but in order to convert the maxi¬ 
mum percentage of the heat energy of the charge into useful 
work the spark must occur a moment before the piston reaches 
the end of the stroke. The result of this is that toward the end 
of the compression stroke the compression curve assumes a 
rising inflection, as indicated in Fig. 8, and instead of joining 
the vertical ignition line sharply, it does so with a curve of con¬ 
siderable radius. 

Another thing which modifies the form of the diagram is the 
fact that the expansion cannot be continued in the cylinder till 
the end of the expansion stroke and the exhaust valve remain 
closed. It must be remembered that the exhaust valve does not 



open instantly but gradually, and if it began to open only at the 
end of the stroke, the piston upon the beginning of the exhaust 
stroke would encounter a back pressure practically equal to the 
pressure which was acting upon it just before the completion of 
the power stroke. There would be a very considerable back 
pressure during a large portion of the return stroke, and this- 
















32 


THE OTTO CYCLE. 


would not only cause a very considerable loss in power, but, 
owing to the retention of the hot gases in the cylinder, would 
tend to overheat the cylinder. For this reason the ex¬ 
haust valve begins to open when about seven-eighths of the 
power stroke is completed, and the pressure in the cylinder then 
drops more rapidly than it otherwise would. This takes the upper 
right-hand corner off the diagram, as shown by line D E, Fig. 8. 

So far we have considered only the curves traced by the pencil 
of the indicator during the compression and expansion strokes. 
When the piston begins the exhaust stroke, and, indeed, through¬ 
out that stroke, the pressure in the cylinder is somewhat above 
atmospheric and the indicator pencil will trace a line, slightly 
curved, a little above the atmospheric line. When the motor runs 
at relatively low speed the exhaust line will join the atmospheric 
line at the end of the stroke, but at high speed it will then still 
be above the atmospheric line, and for this reason the exhaust 
valves in high-speed motors are held open a little longer than 
the end of the exhaust stroke. During the first part of the fol¬ 
lowing inlet stroke the pressure in the cylinder drops fairly rap¬ 
idly, partly because some more of the burned gases escape and 
partly because the space occupied by the gases rapidly increases. 
A short distance from the beginning of the stroke the pressure 
drops below atmospheric, and from that point on to the end of 
the stroke there is a suction or a partial vacuum in the cylinder, 
the suction line lying below the atmospheric line and geing gen¬ 
erally slightly concaved toward the atmospheric line. At high 
speed the pressure is still below atmospheric at the end of the 
suction stroke, and the inlet valve of a high-speed motor remains 
open for a short time after the piston begins the compression 
stroke, because the remaining difference in pressure inside and 
outside the cylinder, together with the inertia of the inrushing 
column of charge insures a continued inflow. From the begin¬ 
ning of the compression stroke the pressure in the cylinder rises. 
The pressure line soon cuts the atmospheric line, and then, as the 
compression proceeds, rises at a constantly increasing rate, as 
already explained. 

When the motor runs at relatively low speed the pressure in 
the cylinder is atmospheric at the beginning and end of the suc¬ 
tion stroke. The pressures during the exhaust and suction 
stroke are so small, as compared with the pressure of explosion, 
that it is necessary to draw this part of the diagram to a different 
scale from the compression and expansion curves, in order to 
show it plainly. 


THE OTTO CYCLE. 


33 


Thermal Efficiency—The thermal efficiency is the ratio of the 
thermal energy actually utilized in the motor to the total amount 
of heat energy supplied to the gases. In the form of an equa¬ 
tion this may be written: 

Energy supplied — energy discharged 

T) =- ; - 

energy supplied. 

Referring to Fig. 9, let 

Ti = Absolute temperature of the gas just previous to ignition; 

T 2 = Absolute temperature of the gas after ignition; 
r 8 = Absolute temperature at the completion of the expansion 
stroke; 

T* = Absolute temperature of the gas as it is drawn into the 
cylinder. 

The heat energy being supplied to the gas at constant vol¬ 
ume, its value—considering a unit quantity of gas—is evidently 
Cy(T 2 — 7 i). The energy is also discharged at constant volume 
at the end of the expansion stroke, the amount so discharged 
being Cy ( T % — T 0 ). We may write, therefore, 


1 


T s -T 0 

n— t x 


_ (T 2 —T x ) Cy — (T,— T 0 ) Cs _ 
( T 2 -T x )Cy 


(13) 













34 


THE OTTO CYCLE. 


But since both the compression and expansion are adiabatic, for 
which form of compression or expansion 


T V*Y 1 = Constant , 

I±-h- (S\ 

T t ~ \vj 


we may write 


hence 


— i 


1 i 


which, substituted in (13), gives 

ZL 

T x 


7 0 - 7 0 


V = 1 — 


T t -T x 

Dividing both the numerator and denominator of the fraction by 

Tt. 

T x 


1 : 


7 ]= I—Zp = 


(-;) ' 


Since in gasoline engine work 7 is always in the neighborhood 
of 1.3, we may finally write 
•3 

.(i 4 N 


— © 


which simple equation gives the theoretical thermal efficiency of 
the Otto cycle gas engine. This equation is of interest in that it 
strikingly shows the influence of the compression ratio on the 
thermal efficiency. Thus, for instance, with a compression ratio 
of 3 the theoretical efficiency is 

I_ (i) M=0 - 281 * 

and with a compression ratio of 4 




34i 


and with a compression ratio of 5 

\ 0-3 


1 — 


= 0.383. 








THE OTTO CYCLE. 


35 


Application of Formula—We will now work out a practical 
example with the aid of the formula developed in the foregoing. 
We will assume an engine cylinder with a bore b = 4 inches and 
a stroke / = 5 inches. Let the compression ratio r- 4. The nor¬ 
mal speed of such a motor is generally about 1,200 revolutions 
per minute, which corresponds to 1,000 feet piston speed per 
minute. It has been found that at such a speed an engine with 
large-sized valves properly timed draws in about 80 per cent, of 
its displacement volume of air per cycle. The air may be con¬ 
sidered at 62 degrees Fahrenheit (=523 degrees absolute). 

The compression pressure will be: 

12 X 4 1 " 3 = 72.7 

pounds per square inch absolute, and the compression temperature 

523 X 4 °' 3 = 792 

degrees absolute, or 331 degrees Fahrenheit. 

The ratio in which the pressure is multiplied on explosion 
varies \^ith different shapes of combustion chamber and also with 
the compression ratio, but may be taken as 3.75 in this case. The 
explosion pressure then will be 

375 x 72.7 = 272.5 

pounds per square inch absolute, and the explosion temperature 

375 x 792 = 2,970 

degrees absolute. 

The ratio in which the absolute temperature is multiplied is, of 
course, the same as the ratio in which the absolute pressure is 
multiplied, since the heat is imparted to the gas at constant 
volume. 

The work of expansion, according to equation (12), is 


IVe = 


3.1416 X 4 2 X3-75 X 72.7 5 


I 4-4 


X ~ X (1 — 0.642) = 568 foot-pounds 


The work of compression, according to equation (10), is 


3.1416 X 4 2 X 12 20 

IV C = - X ~~ X 0.515 = 144 foot-pounas. 

1 4 * 4 3 


The useful work per c^cle is therefore 


568 — 144 = 424 foot-pounds. 

Since at 1,200 revolutions per minute there are 600 explosions 
per minute, the work done per minute is 

600 X 424 = 254,400 foot-pounds. 

Now, one mechanical horse power is equal to 33,000 foot-pounds 
per minute, so this 4x5-inch cylinder develops 




36 


THE OTTO CYCLE. 


2 54 > 4 °° — y ,rj 1 indicated horse 'power. 

33,000 

If 25 per cent, is allowed for the loss due to suction, exhaust 
back pressure and mechanical friction, we get for the net output 
0.75 X 7.71 = 5.8 brake horse power. 

This is just about what would be expected from a motor of 
this size using the rather moderate compression ratio of 4. 

The piston displacement is 

4 X 4 X 5 X 3.1416 _ 62.83 cubic inches. 

4 

and since 80 per cent, of this volume of air is drawn in at each 
suction, the amount of air thus drawn in is 

0.80 X 62.83 = 50.26 cubic inches. 

The amount of air consumed per hour is 


60 X 600 X 50.26 ,. , 

-- = 1,047 cubic feet, 

which weighs 

1,047 X 0.076 = 79.57 pounds, 

and is mixed (if we assume the proportion of air to gasoline to 
be 18 to 1) with 


= 4.42 pounds gasoline. 

Suppose that the gasoline is of 0.72 specific gravity. Then, 
since one gallon of water weighs 8.2 pounds, and since the engine 
develops 5.8 brake horse power, the gasoline consumption per 
horse power-hour is 


&2 X 4 o 72 X 5 . 8 = 0 - 129 ga,lon • 
or about a pint. 

The assumptions made in the foregoing example are in close 
accord with the conditions in an engine of good design, and the 
results give a good idea of the quantities involved. 







CHAPTER IV. 


CONVERSION OF RECIPROCATING INTO ROTARY 

MOTION. 

Number of Cylinders —For small and moderate powers a 
single-cylinder motor offers several important advantages.** It is 
of the simplest possible construction, inexpensive to manufacture 
and more economical in the use of fuel than any other type. 
Along with its advantages it possesses, however, two inherent 
defects, particularly from the standpoint of its use in automobile 
work, for which reason it is now seldom employed on automo¬ 
biles. 

It has already been pointed out that in a four-stroke motor 
only one stroke in every four is a power stroke; between suc¬ 
ceeding power strokes there are three so-called idle strokes, and 
during one of these, the compression stroke, power is even con¬ 
sumed by the motor. In order, therefore, to keep the engine 
running at a fairly uniform speed against a constant or nearly 
constant resistance, it is necessary to employ a heavy flywheel in 
which some of the energy liberated is stored during the expan¬ 
sion stroke, to be given out again during the three following 
idle strokes. But a heavy flywheel is not desirable on a car, be¬ 
cause of the additional weight, which consumes fuel, lessens the 
speed and hastens the wear of the tires. Another objectionable 
feature of the single cylinder motor is its lack of balance. Vi¬ 
brations in a motor are due to the reactions of the explosion and 
of the inertia forces. Now, if we consider a motor of a given 
output, the force of individual explosions is necessarily much 
greater in a single-cylinder motor than in a multi-cylinder motor. 
Moreover, in a single-cylinder motor the entire reciprocating 
mass is in a single unit, and the reactions of its inertia forces 
produce a strong vibrating effect, while in a multicylinder motor 
the reciprocating mass can be divided into several units so ar¬ 
ranged as to move in opposite directions at any given moment, 
thereby neutralizing the inertia effects. 


37 



38 CONVERSION OF RECIPROCATING MOTION. 

The reciprocating motion of the piston is converted into rotary 
motion of the crankshaft by means of an intermediate connecting 
rod. The connecting rod is hinged to both the piston and the 
crankshaft. Its upper end reciprocates in harmony with the 




piston, while its lower end rotates with the crank pin. Before 
proceeding with a discussion of crank and piston motion it may 
be well to state a few fundamental principles which apply to 
these problems. 

Fundamental Relations—Speed is the quotient of the dis¬ 
tance traveled by the time in which it is traveled. If the speed 
is constantly varying, its momentary value can be found by 
dividing the distance d x traveled in an infinitesimal time d t 
by d t. That is, the speed or velocity 
























CONVERSION OF RECIPROCATING MOTION. 39 


V = 


d x 
~dt 


Acceleration is the quotient of a velocity 
by the time in which it was acquired. 
If the acceleration is not constant it 
is necessary in determining its value to 
take an infinitesimal increase in velocity 
d v and divide it by the time d t in which 
this increase in velocity accrued. Hence 
acceleration 

d v d 2 x 




dt d t 2 

The acceleration of a body requires 
the expenditure of energy. A body 
in motion has stored up in it a certain 
amount of energy, called its kinetic energy, 
or vis viva, and this energy is, of course, 
imparted to it when it is set in motion. 

Gravity, which acts on a body with a force 
equal to its weight, imparts a velocity of 
32.16 feet per second to it in one second. 

The unit of velocity is one foot per second, 
though a unit of one foot per minute is also 
used. The unit of acceleration is one foot 
per second per second. The unit of force 
generally employed in mechanical work in 
this country is the pound. Then, since a unit 
of force acting on a unit mass (one pound) accelerates it 32.16 
feet per second per second, and since the acceleration is directly 
proportional to the acting force and inversely proportional to 
the mass acted upon, we may write 



Fig. 12. 


32.16 F / x „ W a 

a = -..(15) or F= - 

IV 32.16 


(16 


The factor 32.16 is known as the acceleration of gravity, and is 
usually denoted by g. 

Piston Motion —In studying the motions of the different parts 
of an engine, it is generally assumed that the crank pin rotates 
at a uniform speed. This supposition is nearly correct. A typical 
arrangement of the piston, connecting rod and crankshaft is 
shown in Fig. 10, and in Fig. 11 these moving parts are repre¬ 
sented diagrammatically. The heavy line — represents the crank- 

2 

arm and the heavy line n l the connecting rod. The lengths of 
these lines represent the “centre to centre” distances; that is, in 











40 CONVERSION OF RECIPROCATING MOTION 

the case of the crank-arm the distance between the centre of the 
crankshaft journal and the centre of the crank-pin, and in the case 
of the connecting rod the distance between the centre of the 
crank-pin bearing and the centre of the piston-pin bearing. It 
will be seen that in Fig. n the crank-arm has moved from the 
upper dead centre position through an angle 0, and the piston has 
in consequence moved down a distance x from its topmost posi¬ 
tion. It is now desired to find the relation between 9 and x, so 
that we may calculate the position of the piston corresponding to 
any position of the crank, as well as the speed and acceleration 
of the piston corresponding to any crank position and a given 
speed of rotation of the crankshaft. 

Referring now to Fig. n, it is obvious that 

x = — (i — cos 0)4- n l (i — cos 0).(17) 

2 

But 

cos 0 = %/ i — sin 2 <P ..(18) 

and 


n / sin 0 — —sin 0 . 

2 

so that 

. , sin o 

sin 9 —- > 

2 n 

which substituted in (18) gives 


(18 a) 


cos 0 


=/ 


sin 2 0 
4 w 2 


Inserting this value of cos 0 in equation (17) we get 


— cos 0 ) + nf (1 — .(*9) 

This equation enables us to determine the portion of the stroke 
completed while the crank is in any position. Now assume that 
the ratio n of the connecting rod length to length of stroke is 2, 
• and let the length of stroke be unity. Then, after a quarter turn, 
when 0-90°, the fraction of the stroke completed will be 

14-2(1 — */ 1 —*■&) = 0.564. 

It is thus seen that when the crank has turned through one- 
quarter of a revolution, more than half of the stroke is com¬ 
pleted. It follows that when the piston has completed half of its 
down stroke less man one-quarter revolution has- been completed 
by the crank, and since at the end of a half a revolution of the 
crank a full stroke is completed and the crank is supposed to 
rotate at uniform speed, the piston travels faster during the first 













CONVERSION OF RECIPROCATING MOTION. 41 


half of its down stroke than during the last half; similarly it may 
be shown that it travels slower during the first half of the return 
stroke than during the last half. In Fig. 12 the piston is shown 
in the maximum speed position, with the connecting rod and 
crank arm in the corresponding positions for both the down 
stroke and the return stroke. 

Piston Speed—The differentiation of equation (19) with the 
object of deriving an expression for the instantaneous piston 
speed involves some difficulties. In fact, the right-hand member 
of the equation can be differentiated only after being developed 
into a series—a rather complicated proceeding. But the difficulty 
can be got around without serious error by adding to the ex¬ 
pression under the radical the term 

sin 1 6 
_________ • 

64 « 4 

This makes that expression a perfect square, and, taking the 
root, we have 


■/ 


1 + 


sin 


* 9 


sin 


2 9 sin 2 9 




64 n K 4 n 2 8 n 2 

Substituting this in equation (19), we get 


x= — (1 — cos 9 ) -f- 
2 


nil’ll) 

' 8»V 


Differentiating, we find for the momentary piston speed 


.,_dx l . ad 9 l . dO 

v - — = — sin 9 - 1 — sin 9 cos 9 — 

2 d t 4 n d t 


d t 


But since the angular speed of the crank is constant 

_ 9 2 tt N 

dt t 60 


.(20) 


radians per second, N being the number of revolutions per minute 
of the crank. Substituting this value and dividing by 12 to get 
the result in feet, we have 


v — 


7r 


IN 


720 


( sin 9 -(- JL gin 9 cos 9 \ feet fer second .(21) 

2 n J 


From this equation the piston velocity corresponding to any 
crank position may easily be found. When the crank is in the 
top dead-centre position ( 9 - o°), and in the bottom dead centre 
position (9= 180°), and sin 9 = o the speed v of the piston is 
nil. Starting from the top dead centre position the speed of the 
piston increases until it reaches a maximum a little before the 
crank comes to the quarter position. The crank position corre¬ 
sponding to maximum piston speed may be found by placing the 













42 CONVERSION OF RECIPROCATING MOTION. 



Fig. 13.—Piston Speed Plotted Against Piston Travel. 

(.N = 1,200 r. p. m. — 1 = 5 " — n = 2.) 











CONVERSION OF RECIPROCATING MOTION. 43 


first differential coefficient of the expression in parentheses in 
equation (21) equal to zero. This gives 


cos 9 -f- — (cos 2 0 —sin 2 0 ) = o; 
2 n 

2 n cos 0 — sin 2 9 — cos 2 9 ; 

2 n . o a 

- = tarn 9 — 1 ; 

cos 9 


Owing to the expedient resorted to when differentiating the ex¬ 
pression for the momentary position of the piston, this last 
equation is not absolutely correct. The maximum speed posi¬ 
tion corresponds nearly with the position for which 

tan 9—2 n, 

that is, when the crank arm and connecting rod make a right 
angle. The parts are shown in this position in Fig. 12. 


tan 6 = 


n l 

y = 2 n. 


, 2 

The variation of the piston speed during the course of one revo¬ 
lution of the crankshaft is shown graphically in Fig. 13, for a 
motor of 4-inch bore and 5-inch stroke with a 10-inch connecting 
rod and running at 1,200 revolutions per minute. 

Piston Acceleration— To obtain an expression for the piston 
acceleration, we again differentiate equation(21), but before doing 


this we can simplify this equation by substituting s i p 2 9 

2 

equivalent, sin 9 cos 9 . This gives us 

v == 7r - ^ -_. ( sin 9 -f- _L_ sin 2 9 j feet per second . 

720 \ 4 n ) 

Then 


for its 

.. .(22) 


d v 
d t 


7r 


IN 


720 


/ a 1 a\ d 9 

I cos 9 4 - — cos 2 9 j — • 
\ 2 n ) d t 


As before, ^^is equal to 2 77 ^ 


7r 


A r 


d t 


60 


30 


and, substituting, we have 


2 AT 2 / / J \ 

a = —_—— ( cos 9 -(- -cos 2 9 ) feet fer second per second..( 23) 

21,600 \ 2 n j 

The acceleration is a maximum when 0 = 0 , as cos 9 and cos 
2 0 both reach their maximum value of 1. This is the dead-centre 
position. Starting from the top* dead-centre the acceleration de¬ 
creases and it becomes nil at the moment the piston reaches its 
maximum speed. It is, of course, self-evident that as the speed 










44 CONVERSION OF RECIPROCATING MOTION. 


ceases to increase and is about to begin to decrease there must 
be a moment of no acceleration. 

From the general expression for the force of acceleration, 


r Wa 
F= - > 

vve may, by inserting the value of the acceleration found in (23), 
derive the following formula for the force of acceleration of 
engine reciprocating parts 

W 7 T 2 ,V 2 / / o , i \ , , 

F* = — -- ( cos 0 -+--cos 2 0 ) -pounds , 

g 21,600 \ 2 n ) 

which can be simplified to 


F&= 0.0000142 WIN 2 


^cos 0 -f- 


1 

2 n 


cos 2 


■) 


pounds 


(24) 


where W is the weight of the reciprocating parts in pounds, l 
the length of stroke in inches, N the number of revolutions per 
minute and n the ratio of connecting rod “centre to centre” 
length to length of stroke. 

Weight of Reciprocating Parts—The inertia force is op¬ 
posed to the force of explosion and materially reduces the shock 
received by the connecting rod and crank when the explosion 
takes place, when the motor is running at considerable speeds. 
The reciprocating weight W includes the weight of the piston 
and rings, the piston pin and about one-half of the weight of the 
connecting rod. If an engine is to be operated at high speed 
these parts must be made very light. They are made much 
lighter in automobile motors than in slow running stationary en¬ 
gines, and still lighter in racing automobile motors. In average 
automobile practice the reciprocating parts weigh about 0.48 
pound per square inch of piston head area. For racing and simi¬ 
lar motors, in which durability is not such an important con¬ 
sideration, and where special materials can be used, this factor 
can be almost, if not completely, halved. 

It is convenient to express the weight of the reciprocating 
parts in terms of the piston area, since the force of acceleration 
acts in opposition to the expanding gas during the first part 
of the power stroke and adds to it during the last part, and 
the force of expansion, as obtained from the indicator or mano- 
graph diagram, is expressed in pounds per square inch. 

The gas-pressure curve represented in Fig. 15 is a portion 
of an actual manograph diagram taken from a French motor of 
4-inch bore, for which diagram tlie writer is indebted to Joseph 
Tracy, of New York. It shows a maximum pressure of 210 
pounds per square inch. Now, let us suppose that the motor 







CONVERSION OF RECIPROCATING MOTION. 45 



(Assumed that 1 = 5 . inches, N = 1,200 revolutions per minute, ^ = 0.55 
pound per square inch, and connecting rod length = 2/.) 


had a stroke of 5 inches and that its reciprocating parts weighed 
0.55 pound per square inch. It is required to find the force of 
inertia at different points of the stroke and the resultant pres¬ 
sure on the connecting rod. 

It is first necessary to find the distance the piston has traveled 
from the top end of the stroke when the crank has turned 
through different angles from o to 180°. This can be done either 
by the construction of a diagram—in other words, graphically— 
or the values can be calculated by means of equation (19). 
The values thus obtained are given in the second column of 
Table I. The piston speeds corresponding to different crank 
angles are next determined by means of equation (22). 
Then the piston acceleration is determined by means of equation 
(23), and finally the force of acceleration by means of 
equation (24). All of the values obtained are given in 








46 CONVERSION OF RECIPROCATING MOTION. 


Table I.—Acceleration Forces in an Engine of 4-Inch Bore, 5-Inch 
Stroke, Connecting Rod Length Twice the Stroke and Re¬ 
ciprocating Parts Weighing 0.55 Pound Per Square 
Inch of Piston Head Area. 





Piston 

Force 


Piston 

Piston 

Acceler- 

of Acceler- 


Position 

Speed (Feet 

ation (Feet P. 

ation (Pounds 

Degrees. 

(Inches). 

Per Minute). 

Sec. P. Sec.). 

Per Square Inch). 

0 

0 

0 

4,110 

70.29 

10 

0.048 

339 

4,012 

68.60 

20 

0.190 

663 

3,635 

62.15 

30 

0.415 

955 

3,185 

54-46 

40 

0.720 

1 .175 

2,602 

44.49 

50 

1.047 

1,396 

L 97 I 

33-71 

60 

1.485 

L 530 

1,205 

20.61 

7 0 

1-925 

i ,598 

495 

8.46 

80 

2-375 

1,614 

— 200 

— 3-42 

90 

2.815 

L 570 

— 836 

— 14-3 

100 

3-245 

1,480 

— i ,345 

— 23.0 

110 

3-635 

i ,347 

— i ,753 

— 30.0 

120 

3-985 

1,190 

— 2,055 

— 35-2 

130 

4.292 

1,009 

— 2,257 

— 39-3 

140 

4-550 

816 

— 2,376 

— 40.6 

150 

4-745 

615 

— 2,404 

— 42.1 

160 

4.890 

411 

— 2,463 

— 42.2 

170 

4.972 

205 

— 2,466 

— 42.2 

180 

5.000 

0 

— 2,466 

— 42.2 

Table I. 

The inertia 

force curve 

is now drawn 

in in Fig. 15. 


At the beginning of the stroke the reciprocating masses must, 
of course, be set in motion, and they then consume some of the 
energy of the expanding charge. When the crank has turned 
through an angle of about 76 degrees the piston reaches its 
maximum speed; thereafter its speed decreases and it returns 
the energy which it absorbed during the first part of the stroke. 
By combining the gas pressure curve and the inertia force curve 
we obtain the curve of the resultant pressure, as shown in Fig. 
15. It will be noticed that the inertia of the reciprocating parts 
greatly reduces the maximum pressure on the connecting rod. 

In the same way that we plotted in Fig. 15, the resultant pres¬ 
sure in the direction of piston travel for the power stroke, we 
may plot this pressure for the other strokes. In doing this the 
gaseous pressures during the admission and exhaust strokes may 
be neglected, as they are exceedingly small in comparison with 
the inertia forces at normal speed. The pressure of the gas dur¬ 
ing compression must, however, be taken into account. All 
forces which oppose the motion of the piston are considered 
negative and are plotted from the base line downward, 
while all forces which assist the piston in its motion are con- 




CONVERSION OF RECIPROCATING MOTION. 47 


sidered positive and are plotted from the base line up. In 
Fig. 16 is shown such a diagram for a complete cycle, the as¬ 
sumptions being the same as for Fig. 15. Starting with the 

admission stroke, the only force acting is that of inertia, and 

since this is a down stroke, the inertia is exactly the same as 
for the power stroke, as plotted in Fig. 15. Exigencies of 
space made it necessary to use in Fig. 16 a horizontal scale 
half as large as in Fig. 15. 

Next comes the compression stroke, during which both gas 

pressure and inertia forces are acting. Owing to the fact that 

this is an up-stroke the inertia force diagram is the reverse of 
the inertia force diagram for the previous stroke. As for every 
other stroke, the inertia force is at first negative and toward 
the end positive. The gas pressure during the compression 
stroke is negative throughout, and in the diagram is represented 
by a dash-dotted line, while the inertia force for this stroke is 
represented by a dotted line. The resultant of the two forces 
is represented by a full line. 

The next following portion of the diagram is an exact re¬ 
production of Fig. 15, except that the horizontal scale is only 
half as large. During the last stroke, while the cylinder is 
exhausting, only the inertia force is acting, and this is again the 
same as during the compression stroke. The curve shown 
in a heavy full line represents, therefore, the variations of the 
effective pressure in the direction of piston travel throughout 
the four strokes of the cycle. 

In the construction of the inertia curves it is convenient to 
make use of a table of “crank angle factors.” The expression 
for the inertia force, viz.: 

0.0000142 Wl N 2 (cos 6 -f- -?-cos2 0 ) pounds, 

. 2 71 

may be considered as composed of two parts, one including the 
different values of the particular engine and the constant, and 
the other depending entirely upon the position of the crank, 
and upon the ratio of the connecting rod centre to centre length 

to the length of stroke. This latter factor (cos 0 + — cos 

2 n 

2 6 ) we will call the “crank angle factor,” and its values for 
angles from zero to 180 degrees in increments of 10 degrees, 
and for values of the ratio 71 from 1^4 to 2}4, inclusive, in 
increments of %, are given in Table II. A rather peculiar 
feature in connection with the crank angle factor is that for 
values of n above 2 the factor is largest for 0 and 180 0 and 
smaller for all intermediate angles, while for values of n below 


/so A 


48 CONVERSION OF RECIPROCATING MOTION. 



i- /Jdmission -*- Cbmjpressi on -l- +ExjpQ7Z9ioft -—I *—Exhaust -1 

Fig. 16.—Diagram of Resultant Pressure in the Direction of Piston Travel in a 4x5 Inch Cylinder. 





















CONVERSION OF RECIPROCATING MOTION. 49 


2 the factor decreases again as the crank angle 
approaches i8o°; in other words, the inertia 
force reaches its second maximum some time 
before the lower dead centre is reached. 

Crank Moment —From Fig. 16 it can be seen 
that the pressure in the direction of piston travel 
is a maximum at the beginning of the power 
stroke. The turning effort on the crank at that 
moment is nil, however, for the reason that the 
connecting rod and crank arms are then in line 
with each other, and all the pressure is spent 
in producing thrust on the crankshaft bearings. 

The turning moment is the product of a force 
into the perpendicular distance from the line in 
which the force acts, to the centre of rotation, 
and is expressed in pounds-inches or pounds- 
feet. 

The pressure in the direction of piston travel 
is held in balance by two other forces, viz., the 
pressure acting along the connecting rod and 
the reaction to the side thrust against the cylinder wall. Fig. 17 
is a so-called parallelogram of forces, in which these various 
forces are represented. If we denote the pressure in the 




Table 

II.—Crank Angle 

Factor 

of Inertia 

Force. 




1 7 A 

2 

2% 

2'A 

2Vs 

2'A 

0° 

1.286 

1.267 

1.250 

1-235 

1.222 

I . 21 I 

I .20 

10° 

1.254 

1.236 

1.220 

1.206 

1.194 

I .183 

I -173 

20° 

1 • 158 

1.144 

1.1 3 1 

1.120 

1.110 

I . I 01 

1.093 

O 

O 

CO 

1.009 

0.999 

0.991 

0.984 

0-977 

0.971 

O.966 

0 

0 

0.816 

0.812 

0.809 

0.807 

0.805 

0.803 

0.801 

50 ° 

0-593 

0.596 

0-599 

0.602 

0.604 

0.606 

O.608 

6o° 

0-357 

0.367 

0-375 

0.382 

0.389 

0.395 

O.4OO 

70° 

0.124 

0.138 

0.151 

0.162 

0.172 

0.181 

0.189 

80 0 

— •095 

— •077 

— .061 

— .047 

— -035 

— .024 

-.OI4 

90° 

— .286 

— .266 

— .250 

— •235 

— .222 

-.210 

-.200 

100° 

— .422 

— .425 

—409 

— •395 

— •383 

— •372 

-.362 

110° 

— .560 

— -546 

— •533 

— .522 

— .512 

— •503 

— •495 

120° 

— •643 

— •633 

— .625 

— .618 

— .6ri 

-.605 

— .600 

I3O 0 

— .693 

— .689 

— .686 

f 

Os 

CO 

4 - 

— .682 

— .680 

— .678 

I4O 0 

—. 716 

—. 720 

— .723 

—-725 

— .727 

— .729 

— •731 

150 ° 

— .723 

— •733 

— .741 

—.748 

— •755 

-.761 

— .766 

160 0 

— .722 

— •736 

— •749 

—.760 

— •77 0 

— •779 

-.787 

170° 

— .717 

— •734 

— •750 

—.764 

— •77 6 

— .787 

— •797 

180 0 

— .714 

— •734 

— •750 

—-765 

CO 

1^ 

1 

— .790 

— .800 














50 CONVERSION OF RECIPROCATING MOTION. 




direction of piston travel by P p the pressure acting along the 
connecting rod by Pp and the side thrust reaction of the cylinder 
wall by P s , then, obviously, 

P c =-JjL- ...(25) 

COS 0 

and 

P s = Pp tan 0 .(25 a) 

In Figs. 18 and 18A are shown two diagrams of the crank 
and connecting rod with the crank at an angle Q beyond the 
top dead centre, which angle is less than 76° in Fig. 18 and 
more than 76° in Fig. 18A. The lever arm through which the 
force F c , acting along the connecting rod, attacks the crank is 

O C= — cos a. 

2 
















CONVERSION OF RECIPROCATING MOTION. 51 


But 


a = go° — (0 + 0) 

in Fig. 18 and 

a ==— [90° — ( 9 + ^)] 

in Fig. 18A, hence, generally, 

a = ± [90° — ( 0 + 0)] , 

and, since the cosine of an angle is equal to the sine of its 
complement, 


cos a = ± sin (0 + 0)..(26) 

and 


— cos a = ± i_ s i n (0 _|_ 0), 
2 2 


which may be written in the form 


± — (sin 0 cos 0 + cos 0 sin 0 ) 
2 


Multiplying the force by the lever arm through which it 
acts, we get for the turning moment 


P l 

M — -—— y —(sin 0 cos 0 -f- cos 0 sin 0) = 
cos 9 2 


Pv l l • a 1 « sin 0\ 

—( sin 0 -f- cos 0 --— ) 

2 \ cos 9/ 

Referring again to Fig. 18, 


sin 0 = 


sin 0 


2 » 


and 


, / sm 2 ^ _ sin 2 0 

cos 0 = 4/ 1 — - 1 — 2“ 

r 4 8 w* 


(approximately). Consequently, 

sin 0 


sin 0 
cos 0 


2 n 


sin' 


0 


4 « sin 0 
8 w 2 —sin 2 0 


8 ?i‘ 


(27) 


(28) 


Substituting in (27) we get 

.,/= (sin * + 4 ” 5io ^ C °l g-'). 

This equation gives the turning moment per square inch 
of piston head area in pounds-inches. It is usually desired, 
however, to get the total turning moment expressed in pounds- 
feet, and to this end the value for M in equation (28) has to be 

















52 CONVERSION OF RECIPROCATING MOTION. 


. multiplied' by the piston head area and divided by 12. This 
gives for the total crank turning moment 

Mt — ^ / sin 0 -f- 4 n s * n ^ cos -pounds-feet .(29) 

96 \ 8 w* — sin 2 0 ) 

which equation enables us to determine the value of the turn¬ 
ing moment on the crank for any angular position. In this 
equation sin 2 0 is quite small as compared with 8 n 2 , and it 
may be neglected in approximate calculations, in which case 
the equation can be reduced to the simple form 

Mt = T ^ /sin 0 -j- s i p ^ cos ® \ pounds-feet. 

95 \ 2 n / 

The moment is nil for 0 = o degrees and 180 degrees, as in 

both of these cases sin 0 = o. It reaches a maximum toward 
the end of the first quarter revolution. For 0 = 90 degrees, cos 
0 = o and the second part of the term in parentheses vanishes. 
The turning moment on the crank is then equal to the prod¬ 
uct of the pressure on the piston into the length of the crank- 
arm— 


M = 17 ^ ~ pounds-feet. 
96 


If we divide the turning moment by the length of the crank- 
arm we get the tangential effort Ft. Hence 

„ 7 r b 2 Pr> f . n 1 4«sin0cos0\ , , , v 

Ft — -- sin 0 + ^— 7 :- 7~rs ) founds . (30) 

4 \ 8 n l — sin 1 0 / 

For 0 = 90° we have 


Ft p v pounds. 

4 

The motion of the crank pin is then for a moment parallel 
to the motion of the piston and since, owing to the connecting 
rod connection, their distance apart cannot change, they must 
travel at the same speed, and the pressures on them in the 
vertical direction must be the same. 

In Fig. 19 is shown a crank moment diagram for a single 
cylinder motor of 4 inches bore, 5 inches stroke, 10 inches con¬ 
necting rod length, 0.55 pound reciprocating weight per square 
inch of piston head area and indicating a compression pressure 
of about 75 pounds per square inch gauge and an explosion 
pressure of 262 pounds per square inch. 

In drawing a diagram of this kind it is convenient to make use 
of a table of crank angle factors for crank pin turning moment, 
like Table III. This table gives the values of the expression 


sin 0 -f- 


4 n sin 0 cos 0 
8 n 2 — sin 2 0 













CONVERSION OF RECIPROCATING MOTION. 53 


0 


Q 



os 

o 

H 

o 


w 

Q 

$5 

l-t 

.J 

>• 

u 

s 

o 

I—» 

c/> 

OS 

o 


o 

< 


£ 

w 


o 

£ 
I—I- 

£ 

OS 

£> 

H 


c> 


<3 

►-< 


















S )4 CONVERSION OF RECIPROCATING MOTION. 


Table III.—Crank Angle Factor of Crank Moment. 


e 

1 H 

iVs 

2 

0° 

0. 

0. 

0. 

10° 

0.223 

0.220 

0.217 

0 

0 

o .435 

0.428 

0.423 

0 

0 

0.626 

0.617 

0.610 

0 

0 

0.788 

0.778 

0.769 

50 ° 

0.913 

0.900 

0.894 

6 o° 

1.177 

0.988 

0.980 

70 0 

O 

00 

1.031 

1.025 

8 o° 

1-037 

1-033 

1.030 

90° 

1. 

1. 

1. 

IOO° 

0-933 

0.936 

0-939 

I 10° 

0.842 

0.849 

0.857 

120° 

0-735 

0-745 

0-753 

130° 

0.624 

0.629 

0.638 

140° 

0-497 

0.508 

0.517 

150° 

0.377 

0.383 

0.390 

160 0 

0.250 

0.253 

0.261 

170° 

0.125 

0.127 

0.130 

180 0 

0. 

0. 

0. 


- n = - 

2% 

2^ 


2V2 

0. 

0. 

0. 

0. 

0.214 

0.212 

0.210 

0.208 

0.418 

0.414 

0:410 

0.407 

0.603 

0.598 

O.592 

0.588 

0.761 

0.752 

O.748 

0.743 

0.886 

0.879 

O.873 

0.868 

0-973 

0.966 

0.957 

0-955 

1.019 

1.015 

I .Oil 

1.007 

1.028 

1.025 

1.023 

1.021 

1. 

1. 

I. 

1. 

0.942 

0.945 

0.947 

0-949 

0.860 

0.865 

0.869 

0.873 

0-759 

0.766 

0.771 

0.776 

0.646 

0.653 

0.659 

0.665 

0.524 

0.531 

0.537 

0.542 

0.397 

0.402 

0.408 

0.413 

0.266 

0.270 

0.274 

0.277 

0.133 

0.135 

0.137 

0.139 

0. 

0. 

0. 

0. 


of equation (29). From a diagram similar to Fig. 16, we obtain 
the piston pressure P p for any crank angle. This we multi¬ 
ply by the area of the piston head 

4 X 7 r = 12.56 square inches, 

and then multiply by the value of —,viz., 2.5 inches, and 

2 

divide the result by 12 in order to obtain the final result in 
pounds-feet. The result obtained is multiplied by the crank 
angle factor obtained from Table III, and the values thus ar¬ 
rived at are plotted as ordinates in a co-ordinate diagram with 
the crank positions as the abscissas. It will be observed that, 
starting from the beginning of the admission stroke, the crank 
turning moment is negative until the crank has turned through 
about 76°; then it becomes positive and remains so until the 
end of the stroke. During the first part of the compression 
stroke for about 107 degrees of crank motion the turning mo¬ 
ment is negative again; then it becomes positive, and just before 
the end of the compression stroke, as the compression pressure 
overpowers the inertia force, it becomes momentarily negative 
again. During the whole of the power stroke the turning mo¬ 
ment is positive. At about the middle of the power stroke 
there is a peculiar dip in the curve. This is due to the fact 
that as the crank approaches the 90° position there is a con¬ 
siderable increase in the piston pressure, owing to the reversal 






zoo 


CONVERSION OF RECIPROCATING MOTION, 


55 



- spUVOc/ 


ft 


S 

< 

os 

c 

< 
>—t 

Q 

H 

55 

W 

S 

o 

s 


M 

55 

< 


os 

w 

Q 

£ 

►J 

>- 

u 


o 

£ 


o 

HH 



























56 CONVERSION OF RECIPROCATING MOTION. 

of the inertia force, while the gas pressure and crank angle are 
nearly constant. During the last stroke, while the engine is 
exhausting, only inertia forces are acting, and the crank pin 
turning moment is negative during the first part and positive 
during the last part. 

Fig. 20 is a crank moment diagram for a two cylinder motor of 
either the opposed type or of the twin type with both pistons 
working on a single crank pin. In both of these engine types 
the explosions occur at equal intervals of one revolution, and 
both of the pistons start simultaneously on corresponding strokes. 
The curve of Fig. 20 is based on the assumption that both cylin¬ 
ders are of 4 inch bore and 5 inch stroke, and the other assump¬ 
tions are the same as in the previous example. In such two cyl¬ 
inder motors the power strokes are separated by only a single 
idle stroke. It will be seen that the maximum turning moment 
is only a little greater than in the case of a single cylinder motor, 
although the power and the average turning moment are twice as 
great. The maximum negative turning moment is, however, also 
about twice as great in the two cylinder as in the single cylinder 
motor. The best criterion of the relative uniformity of turning 
moment is undoubtedly the ratio of the total turning moment 
range to the average turning moment. This is about 9 in the 
case of the single cylinder motor and about 6.4 in the case of the 
double cylinder motor, as may be ascertained from the respective 
diagrams. 



Fig. 21.—Four-Cylinder Crank Moment Diagram. 


<r> 


* 


"Vo 

I 




£ 











CONVERSION OF RECIPROCATING MOTION. 57 



Fig. 22. —Six Cylinder Crank Moment Diagram. 


Fig. 21 is the turning moment diagram of a four cylinder motor 
of 4 inches bore and 5 inches stroke. Under the assumptions 
here made the turning moment is always positive, starting, of 
course, from o at the beginning of each stroke. It is almost nil 
at 50° past the top dead centre, and it would become negative at 
this point if the reciprocating parts were heavier or the gas pres¬ 
sure relatively less. The ratio of the total turning moment range 
to the average turning moment is 2.4. This diagram very strik¬ 
ingly shows the effect of the inertia of the reciprocating parts. 
Although the gas pressure in the cylinder is far greater during 
the first half of the down stroke than during the last half, the 
turning moment on the crank is almost nil during the first half 
of the stroke, and attains its maximum value during the last 
half, at about 125 degrees from the top dead centre. 

Fig. 22 is the turning moment diagram of a six cylinder 4x5 
inch engine. The ratio of the maximum turning moment range 
to the average turning moment is 1. Of course these different 
factors of turning moment fluctuation, viz., 9 for the single cylin¬ 
der, 6.4 for the double cylinder, 2.4 for the four cylinder, and 1 
for the six cylinder motor, apply strictly only if the assumptions 
here made regarding gas pressure variation, weight of recipro¬ 
cating parts and relative length of connecting rod hold true, but 
the factors would be very nearly the same for any practical au¬ 
tomobile motor, and the above figures may well be taken as in¬ 
dices of the relative uniformity of turning moment in motors 








58 CONVERSION OF RECIPROCATING MOTION, 



Fig. 22A. —Eight-Cylinder Crank Moment Diagram. 


with various numbers of cylinders. The four cylinder and six 
cylinder crank moment diagrams are shown only for one revolu¬ 
tion of the crankshaft, which is all that is necessary, since the 
four cylinder diagram repeats itself after each 180 degrees of 
crank motion and the six cylinder after each 120 degrees. 

A turning moment diagram for an eight cylinder V type motor 
is shown in Fig. 22A. The ratio of maximum variation in turning 
moment to average turning moment in this case is about 0.8, thus 
showing a further improvement over the six cylinder engine. 
The eight cylinder compares even more favorably with the six in 
this respect when low speed turning moments are considered. 

The crank moment diagrams may also be made to serve as 
crank effort diagrams by suitably dividing the vertical scale. 
Since the crank arm in all the foregoing examples is 2 x / 2 inches = 
0.208 ft., the tangential effort in pounds on the crank pin for any 
angular^position of the crank is equal to the corresponding crank 
moment in pounds-feet divided by 0.208. In Figs. 21 and 22 the 
crank effort scale is drawn in at the right of the diagram. 






CHAPTER V. 


BALANCING OF ENGINES. 

There are three possible causes of vibration in a gasoline 
motor, viz.: 

(1) Unbalanced rotating parts. 

(2) Unbalanced reciprocating parts. 

(3) Torque reaction. 

Aside from the flywheel, which when properly made is per¬ 
fectly balanced, the only rotating part of considerable inertia 
is the crankshaft. If an ordinary single throw crankshaft is 
rotated at considerable speed, the centrifugal force acting on 
it is transmitted to the engine frame and always tends to dis¬ 
place the engine in the direction in which the crank arms are 
pointing at the moment. Since the crank arms are rotating uni¬ 
formly, the general effect, if the engine were absolutely free, 
would be to displace its crankshaft centre line in a circle. The 
rigid connection of the engine to the chassis frame, as well as 
the more or less flexible connection of the latter to the axles and 
wheels, materially modifies and reduces the amplitude of the re¬ 
sulting vibration, but does not eliminate it entirely. 

Crankshaft Balance —To obviate such vibration the crank¬ 
shaft must be balanced; that is, it must be provided with counter 
weights extending from it radially in the opposite direction from 



Fig. 23.—Crankshaft With Balance Weights. 


V 

\ 


59 
































60 


BALANCING OF ENGINES. 


the crank arms, so proportioned that if the crank is placed on a 
pair of steel balance ways it will remain in any position in 
which it is placed. (See Fig. 23). Cranks with more than one 
throw, in which the throws are symmetrically distributed, are 
inherently balanced if carefully made, with all crank arms of 
exactly like dimensions, except for the fact that the radial cen¬ 
tre lines of the arms, etc., on one side of the crankshaft may not 
coincide with the radial centre line of the arms on the other 
side. This occurs in the double throw crankshaft of two cylinder 
motors, and produces a slight rotating couple. In most practical 
cases this can be neglected, but if great stress is laid upon the 
best possible balance, such cranks can also be provided with 
balance weights or counterweights opposite each short crank 
arm. 

The balancing of the rotating parts presents no particular 
difficulty in any type of engine, being merely a matter of 
care in construction. The reciprocating parts, however, are 
not so easy to deal with. Considering first the case of the single 
cylinder motor, its reciprocating parts might be balanced* by 
outside weights connected to the crankshaft in such a manner 
as to be moved at every instant at the same speed as the engine 
piston but in the opposite direction. There should be two such 
weights, one on either side of the motor, each equal to half the 
weight of the reciprocating parts. Such an arrangement is not 
practical, however, because, rather than complicate a single 
cylinder motor in this way, any designer would prefer to adopt 
a multicylinder motor. 

Reciprocating Masses —It is, however, possible to partly 
balance a reciprocating mass by a revolving mass. This may be 
explained as follows: A revolving mass is acted upon 
by a radial force (centrifugal force) 

F= 1.226 WN 1 r founds .(31) 

where W is the weight in pounds, N the number of revolutions 
per second and r the radius in feet. Such a rotating force can 
be decomposed into two forces at right angles to each other. 
We will decompose it into a vertical and a horizontal component. 
Referring to Fig. 24, let O A represent the force in magnitude 
and momentary direction. Then the vertical component 

O B = 0 A cos 9 
and the horizontal component 

BA=OAsin6 

When 0 = o, that is when the weight is directly above the centre 



BALANCING OF ENGINES. 


61 



of rotation, the vertical component is equal to the total force and 
the horizontal component is nil. The same when the weight is 
directly below the centre of rotation, when 0 =i 8 o degrees. Each 
of the components can be represented by a sine curve; that is, a 
curve whose abscissas represent angles and whose ordinates are 
proportional to the sines of the angles represented by the corre¬ 
sponding abscissas. (See Fig. 25.) 

Single Cylinder Motor —Now let us consider again a single 
cylinder 4x5 inch engine with reciprocating parts weighing 0.55 
pound per square inch of piston head area. We found that the 
inertia force was 70 pounds per square inch at the beginning of 
the down stroke and 42 pounds at the end of the down stroke and 
the beginning of the up stroke, giving a mean of 56 pounds per 
square inch or 

12.56x56 = 703 pounds 

for the total reciprocating weight. If we provided the crank with 
a balance weight, such that at 1,200 revolutions per minute the 
centrifugal force on it were equal to 703 pounds, the vertical 
component of this force would substantially balance the inertia 
force of the reciprocating parts, but we would then have an equal 
horizontal component left with nothing to balance it. The net 
result of this arrangement would obviously be to transform the 
vertical reciprocating force into a horizontal one. whereby nothing 
would be gained. But if we provide the crankshaft with such 
a weight that the centrifugal force is only half the inertia force 
of the reciprocating parts, then the latter will be reduced by one- 
half and a horizontal component equal to it will be introduced. 
The maximum unbalanced force is therefore halved, and the 












62 


BALANCING OF ENGINES. 



amplitude of the vi¬ 
brations will also be 
halved. In our ex¬ 
ample, considering the 
balance weight to be 
centralized at crank 
arm length from the 
centre of rotation, the 
necessary weight can 
be calculated from 
equation (31), which 
can be transformed so 
as to read 

W — F 

1.226 N 2 r 


IV = -^-= 6.0 pounds , 

1.226 X 20 X 20 X o. 208 

and one-half of this is 3.45 pounds. 


There is, however, still another method of balancing the re¬ 
ciprocating mass by rotating masses, which is employed in the 
Brush runabout single cylinder motor. Suppose that from the 
crankshaft we drive by spur gearing, so as to revolve in the oppo¬ 
site direction and at the same speed, a countershaft, carrying a 


balance weight of 3.45 pounds, 
the gears are so meshed that 



Balancing System. 


the same as the crankshaft, and that 
both balance weights are vertically 
above their centres of rotation at 
the same time. Then the vertical 
components of the centrifugal 
forces acting on the two balance 
weights are evidently equal to the 
mean of the inertia forces, and the 
two horizontal components of the 
two centrifugal forces neutralize 
each other because the two balance 
weights revolve in opposite direc¬ 
tions.. If it were not for the fact 
that the inertia force is not the 
same for a point in the first 
quarter revolution as for the corre¬ 
sponding point in the second quar¬ 
ter, and for the further fact that 
the vertical component of the cen- 














63 


BALANCING OF ENGINES. 

trifugal force on the countershaft balance weight is not in line 
with the inertia force of the reciprocating parts, the reciprocat¬ 
ing masses of this engine would be perfectly balanced. 

Torque Reaction—The third cause of vibration is the torque 
reaction. When the connecting rod impresses a certain turning 
moment upon the crankshaft, the side thrust of the piston against 
the cylinder wall causes an equal and opposite moment to be 
impressed upon the engine frame. This can easily be shown 
algebraically, but it is self-evident, since, according to the laws 
of mechanics, action and reaction are always equal and opposite. 
The only way in which this cause of vibration can be eliminated 
is by having two equal engines side by side, in a single block 
and turning synchronously in opposite directions by being geared 
together. This solution of the problem was repeatedly suggested 
in the early years of automobile history, but was never applied to 
any extent. It was thought preferable to use multicylinder en¬ 
gines, thus reducing the size of the individual cylinders and the 
power of the individual explosions. The fluctuations in the turn¬ 
ing moment will then be much smaller. It must be remembered 
that what causes vibration is not an unbalanced force but the 
fluctuations of such a force. A steady uniform turning moment 

might twist the motor slightly out of its normal position, but 
the engine would remain in that position as long as the torque 
was maintained and there would be no vibration. But if the 
moment varies periodically, or even changes In direction, then 
the motor will move back and forth, or vibrate. 

Amplitude of Vibration—A vehicle motor may be consid¬ 
ered secured rigidly to the vehicle frame and the latter supported 
freely upon the vehicle springs. That is, the resistance which 
determines the amplitude of the vibration is simply the inertia 
of the forward part of the chassis. We found that in our single 
cylinder 4x5 inch engine there is an upward thrust of 

12.56 X 70 = 880 founds 

on the engine frame due to the inertia of the reciprocating parts. 
As the piston descends this force decreases, but at any point up 
to that of maximum piston speed the engine frame is pulled up 
with the same force with which the piston is pulled down, be¬ 
cause action and reaction are equal and opposite. If we call 
this force F, the weight of the reciprocating parts zv, the weight 
of the forward part of the chassis W, the velocities of the two 
v and V respectively, then 


64 


BALANCING OF ENGINES. 


_ Fg 

7JU 


and 



W 

Hence, 

II 

< 8^5 


and, since the time of motion is the same for both and the 

velocities always bear the same ratio — » it follows that the ratio 

v 

of the distances moved by the reciprocating parts and the engine 
frame 


D zu 

In our example, since the reciprocating parts weigh 

12.56 X 0.55 = 6.9 founds 

and they travel about 2.3 inches before their acceleration ceases, 
if we assume that the forward part of the car weighs 400 pounds, 
then it will be thrown up 

x 2.3 =0.04 

400 

or a little over 1-32 inch. 

It can also be proven that the reaction of the torque M will 
cause a point located a distance r from the centre of the 
crankshaft to be displaced a distance 

, Mrt 2 
d — - 

where t is the time in seconds the torque M acts and I the mo¬ 
ment of inertia. But since probably no one has any idea of 
the magnitude of the moment of inertia of the frame with 
motor and body, around the crank shaft centre line, this formula 
is merely of academic interest. 

Two Cylinder Motor—Considering next the two cylinder 
motor, the only type that is now being used to any extent for 
automobile work is that in which the cylinders are located hori¬ 
zontally on opposite sides of the crankshaft, the latter having 
double throws.** Although such motors have been made with two 
cylinders exactly in line with each other, generally they are 
slightly displaced sidewise with relation to each other. Since 
the two pistons start simultaneously from the outer end of the 
stroke (which corresponds to the top end in a vertical motor) 




BALANCING OF ENGINES. 


65 


their velocities are always equal, and since the two sets of re¬ 
ciprocating parts are equal in weight their inertia effects would 
neutralize each other if it were not for the slight sidewise dis¬ 
placement of the two cylinders. As it is, the forces form a 
couple in a horizontal plane, proportional to the distance be¬ 
tween the cylinder centre lines. The moving parts in such a 
motor are thus very nearly balanced, and since the range in 
the torque reaction for a motor of a given output is much smaller 
than in the single cylinder motor, especially at low speed, a 
double cylinder opposed motor runs relatively free from vi¬ 
bration. 

Two cylinder vertical motors have been used to quite an ex¬ 
tent in the past, but have lately’ been entirely discarded. There 
are two possible arrangements. The two pistons may either 
work on a single crank, in which case the explosions in the 
two cylinders will be equally spaced, but there will be a large 
unbalanced inertia force, since the two pistons move up and down 
together; or, the two pistons may act on cranks set opposite or at 
i8o°, in which case one explosion will follow another after one 
stroke and then there will be an interval corresponding to three 
strokes before another explosion occurs. In the latter con¬ 
struction the reciprocating masses will be partly balanced, since 
one set moves up while the other moves down. But the in¬ 
stantaneous velocities will, of course, always be different; and, 
besides, the two reciprocating masses are located at a considerable 
distance apart and produce, therefore, a rather strong couple. 

V Type Motor.—Two cylinder motors with the cylinders set 
at an angle of 90 degrees (and sometimes less) are used to quite 
an extent on motorcycles and the lightest types of four wheeled 
vehicles. They are known as V type motors and have a rela¬ 
tively good balance, considering the small number of cylinders. 
In such a motor the masses reciprocate along lines at right angles 
to each other. When one piston begins its stroke the other is 
nearly at mid-stroke; hence, when the inertia force due to the 
reciprocating mass of one cylinder is a maximum, that due to the 
reciprocating mass of the other cylinder is nearly zero. We 
found that the inertia force on the reciprocating masses varies 
as the expression 

cos 0 4- —— cos 2 0 

2 n 

The first part of this expression, which is by far the most im¬ 
portant, may be represented by a sine curve. But two sine curve 


66 BALANCING OF ENGINES. 

forces acting at right angles to each other, and so timed that one 
is zero when the other is a maximum, are equivalent to a radial 
rotating force, and we can therefore balance the major portion 
of the inertia forces by means of a suitable balance weight placed 
on the crankshaft opposite the crank arms as shown in Fig. 26. 
The balance is the more nearly complete the greater the ratio of 
connecting rod length to stroke and would be complete for an 
infinitely long connecting rod. 

Four Cylinder Motor—In a four cylinder vertical motor (the 
type now most extensively used on automobiles) the two inside 
pistons work on cranks in the same plane, and the two outside pis¬ 
tons on cranks directly opposed to the former (Fig. 28). The 
two inside pistons then always move up and down together, and 



Fig. 28.—Sketch of Four 
Cylinder Engine. 


Fig. 29.—Sketch of Six 
Cylinder Engine. 


the two outside pistons at all times move in the opposite direction 
to the inside pistons. The centre lines of motion of the two sets 
of reciprocating masses coincide. However, although the two sets 
of reciprocating masses are equal in weight and always move in 
opposite directions, the inertia forces do not completely neutralize 
each other, for the reason that, as already explained, the pistons 
travel slower when the crank is a certain angular distance from 
the lower dead centre than when it is the same angular distance 
from the top dead centre. 

In Fig. 30 are shown the reactions on the engine frame due to 
the inertia forces of the reciprocating masses. The inertia forces 
on the two sets of reciprocating masses are shown in dotted 
lines, and the resultant of the two is shown in a full line. Up¬ 
ward reactions on the engine frame are plotted above the hori- 































BALANCING OF ENGINES. 


67 



Fig. 30.—Four Cylinder Inertia Force Diagram. 

zontal axis and downward reactions below the horizontal axis. It 
will be noticed that under the conditions assumed here the unbal¬ 
anced inertia force is about one-half what it would be in a double 
cylinder motor of the same bore and stroke, which, however, 
would have only half the output. 

The turning moment in a four cylinder motor is (as has al¬ 
ready been shown) very much more nearly uniform than that of 
a two cylinder motor, hence the torque reaction and the vibration 



F IG . 31.—Six-Cylinder Inertia Force Diagram. 














68 


BALANCING OB' ENGINES. 


due to it are very much smaller. On the whole the four cylinder 
motor, when properly constructed, meets the requirement of vi¬ 
brationless running quite satisfactorily. 

Six Cylinder Motor—The most nearly perfect motor, from 
the standpoint of balanced running, is the six cylinder, in which 
the pistons act on cranks set two and two together at angles of 
120°, pistons Nos. i and 6 connecting to cranks in the same 
plane, also pistons Nos. 2 and 5 and pistons Nos. 3 and 4 (Fig. 
29). The reactions of the inertia forces due to the reciprocating 
masses of such a motor are graphically shown in Fig. 31, and it 
will be seen that the resultant of all of these inertia forces is zero 
for all crank positions. That is, the reciprocating masses are in 
absolute balance. The variation in the turning moment, and con¬ 
sequently in the torque reaction, is also much smaller than in any 
engine with a smaller number of cylinders, the total range of turn¬ 
ing moment in the example cited being equal to the average turn¬ 
ing moment in the case of the six cylinder motor, while in the 
four cylinder motor it was 2.4 times the average turning moment. 
Consequently, a six cylinder motor may be made to run exceed¬ 
ingly smoothly. To insure this excellent balance it is, of course, 
necessary that all reciprocating parts be made of standard weight; 
that is, each piston must weigh the same as every other piston, 
and the same with respect to the connecting rods. 

Eight Cylinder Motor—An eight cylinder V type motor con¬ 
sists essentially of two four cylinder motors placed at right 
angles to each other (the usual arrangement). Each of these 
motors has an unbalanced force in the central plane of its 
cylinders, as represented by Fig. 30. Now, a force in one plane 
cannot possibly balance a force acting in a plane at right angles 

thereto. The resultant 
of the two unbalanced 
forces is in a horizontal 
plane, as shown in Fig. 
31 A. The maximum 
value of the unbalanced 
inertia force of an eight 
cylinder motor is 1.42 
times that of a four cyl¬ 
inder motor of the same 
cylinder dimensions, but, 
since the eight-cylinder 
motor has almost twice 
the weight of the four, 
i t s relative balance i s 
somewhat better than 
that of the latter. 



Fig. 3iA.—Inertia Force in Eight- 
Cylinder V Engine. 

A—Unbalanced force of right-hand set of 
cylinders. 

B—Unbalanced force of left-hand set of 
. cylinders. 

C—Resultant unbalanced force. 





CHAPTER VI. 


THE CYLINDER. 

Horizontal versus Vertical Cylinders—in the early years 
of the automobile industry in this country, when single cylin¬ 
der motors were used almost exclusively, the cylinders were 
generally placed horizontally. Later on, when the number of 
cylinders was increased, the vertical arrangement was adopted, 
because a vertical multicylinder motor requires less room on 
the chassis and can be arranged more accessibly and more 
symmetrically. It also possesses certain advantages with re- 
. ^pect to lubrication. The horizontal arrangement is retained, 
however, in the case of the double cylinder motor, in which 
the cylinders are placed on opposite sides of the crankshaft, 
on account of the good balance thus obtained. It was often 
urged against the horizontal motor that the weight of the 
piston had a tendency to wear the cylinder bore oval, but it 
can easily be shown that in comparison with the side thrust 
of the piston on the cylinder wall due to the angularity of the 
connecting rod, the piston weight is absolutely negligible, and 
this feature had nothing to do with the abandonment of the 
horizontal motor. 

Arrangement of Valves—The cylinder of a gasoline motor 
may be given many different forms, the differences depending 
chiefly upon the location of the valves. In designing a cylin¬ 
der or group of cylinders, there are two main factors to be 
kept in view, viz., the form of the combustion chamber and 
the means of actuating the valves from the crankshaft. The 
combustion chamber should have the least wall surface in pro¬ 
portion to its cubical contents, as the loss of heat through 
the walls depends very largely upon the extent of this surface. 
This consideration would lead to the adoption of a spherical 
combustion chamber, for of all geometrical forms the sphere 
nas the least surface area for a given volume. This ideal form 
of combustion chamber is approached in the Knight sleeve valve 
motor, as built by the Daimler Motor Company, of Coventry. 


69 



70 


THE CYLINDER. 


England. With poppet valves the nearest approach to an ideal 
combustion chamber is obtained when both inlet and exhaust 
valves are located in the cylinder head at an angle of 30 to 45 
degrees, the cylinder head being made of hemispherical form, 
or domed. The piston head can, of course, also be hollowed 
out, as it is in the Daimler-Knight motor, but this is seldom 
done. The next best arrangement from this point of view is 
that denoted by 3 in Fig. 32, where both valves are located 
in the cylinder head side by side and parallel. This gives a 
cylindrical form of combustion chamber. Next follow the ar¬ 
rangements 2, 4 and 6, which are about equal. In arrange¬ 
ment 2 the inlet valve is located directly above the exhaust 
valve in a side pocket and opens downwardly. In arrange¬ 
ment 4 the inlet valve is located in the centre of the cylinder 
head, and the exhaust valve in a side pocket, while in arrange¬ 
ment 6 the two valves are located opposite and in line with 
each other in a valve pocket formed on the cylinder head. In 
each of these cases there is a valve pocket the =ire of which 
is determined by the size of one valve head. These pockets 
naturally add considerably to the wall surface for a given 
combustion space volume. Next in order of combustion space 
wall area comes the arrangement 7, with both valves in a 
single pocket on the side of the cylinder, and finally that de¬ 
noted by I; with the inlet and exhaust valves in side pockets 
on opposite sides. 

Minimum wall area is, however, only one consideration in 
the design of a cylinder, and, in the estimation of many de¬ 
signers, by no means the most important. Upon the arrange¬ 
ment of the mechanism for actuating the valves depends to a 
large extent the more or less noiseless operation of » motor 
as well as its general appearance, and, to some extent, its 
ability to run at high speeds and to give a large output in pro¬ 
portion to its dimensions. This mechanism is of the simplest 
form when both valves are located in a single side pocket, in 
which case only a single camshaft for operating the valves, 
and a single set of camshaft gears are needed. The cam¬ 
shaft is placed directly underneath the valve chamber and the 
valves are lifted by direct thrust. With the valves in pockets 
on opposite sides the method of operation is similar, but two 
camshafts and two sets of camshafts gears are required. When 
the valves are arranged in an inverted position, as when 
they are located in the cylinder head or when the inlet vaive 
is located above the exhaust valve, a more complicated inter¬ 
mediate mechanism is required, with more chances for devel- 


THE CYLINDER. 


71 


oping play at the joints. Sometimes in multicylinder motors 
the camshaft extends centrally across the tops of the cylin¬ 
ders, being driven from the crankshaft through an interme¬ 
diate vertical shaft or through a silent chain, but more fre¬ 
quently there are a series of vertical valve rods at the sides of 
the cylinder, which are moved upward by the cams, and the 
direction of motion is then changed by short tappet levers 
whose arms bear against the valve stems and the valve rods 
respectively. From the viewpoint of simplicity of valve gear, 
arrangement 7 is undoubtedly the most advantageous. The 
T head motor (arrangement i), on the other hand, permits of 
using larger valves. 

Another question that affects the problem of valve location 



Fig. 32.—Valve Arrangements. 


is that of effective cooling of the valve seats. Practically all 
cylinders are now made with their heads integral, and if the 
valves were to be located in the head and the seats for the 
valve formed in the metal of the cylinder casting, whenever it 
was necessary to remove a valve, or to grind it in, the entire 
cylinder or group of cylinders would have to come off the 
base. This is obviously impracticable, and the necessity for it 
is obviated by inserting cylindrical shells, known as valve 
cages, into the cylinder heads, on which the valve seats are 
formed. Unfortunately, when the valve seat is on an inserted 
cage it cannot be cooled as effectively as when it is in the 
cylinder casting itself, for the simple reason that the heat in 
traveling from the valve seat to the water jacket must pass a 
joint between two metal parts. The reduced cooling power is 












T2 


THE CYLINDER. 


of no consequence in the case of the inlet valve, which is kept 
cool by the fresh charge passing through it, but the exhaust 
valve will heat more and has to be reground more frequently 
when operating in a cage than when seated on the cylinder 
casting directly. 

A cylinder with valves in the head will give more power 
than a cylinder with valves in side pockets, of the same di¬ 
mensions. According to Enrico Giovanni, chief engineer of 
the Fiat Automobile Works of Turin, Italy, the valve-in-the- 
head motor gives as much as 20 per cent, more power than a 
T-head motor of the same bore and stroke. Nevertheless, 
for touring car motors he prefers the T-head type; for one 
reason, because the spark plugs can be located in the inlet 
valve chamber, where there is always a pure charge, and thus 
ignition can be effected when the motor is greatly throttled. 
In other words, a T-head motor can be run at a lower idling 
speed than a valve-in-the-head motor. The valves can also 
be more easily removed. The valve-in-the-head motor, how¬ 
ever, on account of its greater power output, is now used al¬ 
most exclusively for special racing machines. One other ad¬ 
vantage of this type of motor is that the entire combustion 
chamber can be machined, which permits of accurately equal¬ 
izing the compression volumes in the different cylinders, and 
of polishing the walls of the combustion chamber with a view 
to preventing the accumulation of carbon and consequent pre¬ 
ignition. 

Other considerations that come up in the determination of 
the most suitable valve arrangement are the attachment of the 
admission and exhaust pipes and the possible enclosure of the 
valve mechanism. In weighing the advantages and disad¬ 
vantages of the various arrangements against each other, dif¬ 
ferent designers naturally reach different conclusions. The 
following statistics gathered at recent shows will be of inter¬ 
est in this connection. At the Madison Square Garden, New 
York, Show, Part I (pleasure vehicles), in January, 1911, 40 
per cent, of the motors had T-head cylinders (arrangement 1, 
Fig. 32) ; 38 per cent. L-head cylinders (arrangement 7), and 
12.5 per cent, valve-in-head cylinders (arrangements 3 and 5). 
At the London Olympia Show, December, 1910, counting only 
motors with valves on the same side and valves on oppo¬ 
site sides, 78 per cent, were of the former and 22 per cent, of 
the latter type. 

The valve arrangements shown in Fig. 32 are all suitable 
for vertical motors. With horizontal motors arrangement 7, 


THE CYLINDER. 


rd 

v ith the two valves side by side on top of the cylinder, is 
almost universally used. 

Cylinder Grouping—In motors with more than one cylinder 
the cylinders may be cast either singly, all in one “block.” or 
in two or more dual or triple units. Since no two cylinder 
motors except the opposed type are at present being used in 
automobile work, the question of cylinder grouping arises 
only in connection with four and six cylinder motors. The 
use of separately cast cylinders for “fours” and “sixes” has 
been practically abandoned, except where special conditions, 
such as the use of sheet copper jackets, make their use neces¬ 
sary or specially desirable. The advantages of multicylinder 
blocks over twin cylinders are practically the same as the ad¬ 
vantages of twin over single cylinders. By casting all of the 
cylinders in a single block the over-all length of the motor is 
reduced and the construction generally is simplified. In a 
four cylinder motor, for instance, when separately cast cylin¬ 
ders were used, it was customary to have five supporting bear¬ 
ings for the crankshaft; with twin cast cylinders three crank¬ 
shaft bearings are generally used, and with all four cylinders 
cast in a single block the crankshaft is often supported in 
only two bearings, one at either end. In this way the weight, 
as well as the bulk, of the motor is reduced Another effect 
of casting several cylinders in a single block is that the con¬ 
nections are greatly simplified. In a four cylinder block 
motor the jacket spaces for all four cylinders are in one, 
hence there need be only one cooling water inlet and one 
outlet, against four of each in the case of a motor with sep¬ 
arately cast cylinder^. Moreover, in a “block” motor the 
inlet and exhaust manifolds may be cast integral with the 
cylinders, in which case there is only one inlet and one ex 
haust pipe connection. Such a motor is of exceedingly “clear 
cut” appearance. A further advantage of b’o^k construction 
resides in the fact that a housing for the valve springs may be 
cast integral with the cylinder block, so that by merely fitting 
a plain cover plate the entire valve mechanism .'pay be neatly 
encased. 

The disadvantage of block construction is thai if a single 
lefect occurs in a casting either while it is being machined 
or later in service, the entire block may possibly have to be 
thrown away, instead of only a single cylinder. The b-'oek cyl¬ 
inder is naturally more difficult to cast than the single cylin¬ 
der, but the technique of foundry work has been greatly 
improved in recent years and there are now many foundries turn- 


74 


THE CYLINDER. 


ing out block motors with integral manifolds. Owing to the more 
difficult work and greater loss from “wasters,” founders charge a 
somewhat higher price for block cylinders than for single cylin¬ 
ders, pound for pound. The disadvantages of block construction 
are most apparent when a defect is found in a casting after the 
machining has been completed or if the workman spoils a casting 
at this stage; if a cylinder bore is ruined by failure of the 
lubrication or other cause, or a cylinder wall is cracked by 
freezing of the jacket in cold weather. 

It will be seen that the advantages of block construction 
are fundamental or inherent, while its disadvantages are 
largely dependent upon “the state of the art.” These disad¬ 
vantages vanished as experience accumulated, lubricating methods 
were rendered more reliable and means for repairing injuries to 
cylinders were installed in many repair shops. Therefore, as 
was to be expected, in four cylinder engines of moderate output 
block cylinder construction superseded twin cylinder construc¬ 
tion, the same as twin cylinders at an earlier period superseded 
separately cast cylinders. As regards foreign practice, it may 
be stated that of the four-cylinder motors, seen at the Lon¬ 
don Olympia Show in November, 1910, 52.5 per cent, had 
twin cylinders, 37 per cent, block cylinders and 10.5 per cent, 
separately cast cylinders. At the Paris Show in December, 
1910, the proportions were: 50 per cent, block cylinders, 40 
per cent, twin cylinders and 10 per cent, separately cast cylin¬ 
ders. Block cylinder construction has also become exceedingly 
popular in this country, particularly for four-cylinder motors. 
It is at present applied especially to small engines. A few six- 
cylinder motors have all six of their cylinders cast in a single 
block, but the majority of six cylinder motors have either 
three twin or two triple cylinders. 

Cylinder Material—Automobile engine cylinders are gener¬ 
ally cast from close grained gray iron approximating the fol¬ 


lowing composition: PerCent 

Carbon . 3.25 

Silicon . 2.00 

Phosphorus . 0.75 

Manganese . 0.50 

Sulphur—not to exceed. 0.10 


Such iron has a minimum tensile strength of 24,000 pounds 
per square inch and pours well. The use of steel for cylinders 
has often been suggested, and for racing and flying machine 







THE CYLINDER. 


75 


motors cylinders are occasionally worked out of the solid 
oteel billet. One American manufacturer uses cylinders cast from 
semi-steel. Of late years a material known as vanadium iron has 
come into use to some extent for cylinder castings. Vanadium 
present in small quantity in the iron ore accomplishes the same 
result as when incorporated into steel; that is, it eliminates some 
of the objectionable impurities, acting as a “purger.” 

Following are the compositions of domestic and foreign cylin¬ 
der irons respectively, as required by the specifications of the 
American Locomotive Company (for which I am indebted to 
B. D. Gray, chief engineer of the company) : 


DOMESTIC IRON 

Silicon . 

Sulphur. 

Phosphorus . 

Manganese . 

Combined carbon. 

Graphitic carbon. 


CASTINGS. Pcr Cent 

. I .25 to I .60 

.less than 0.11 

. o.50 to 0.80 

.. 0.50 to 0.80 

.over 0.40 

.under 3.00 


IMPORTED CYLINDER CASTINGS. 


Silicon . 

Sulphur. 

Phosphorus . 

Manganese . 

Combined carbon 
Graphitic carbon. 


. 2.50 to 3.00 

under 0.10 

. 1.50 to 1.80 

.. 0.30 to 0.50 

,.over 0.50 
under 3.00 


Among the other points covered by these specifications are the 
following: Castings must be made from hard close grained iron; 
free from shrinkage cracks, spongy spots, blow-holes and foreign 
substances; true to pattern and must not develop any defects in 
machining. Drillings taken from several castings from each day’s 
cast for analysis must show the above composition. When speci¬ 
fied on the drawing, the maker’s brand or trademark and date of 
pouring must be cast in raised figures on every casting at the 
place designated on the drawing. The order and requisition 
numbers must be painted on each casting with white lead. The 
manufacturer must furnish two test pieces i 1 /^ inches in diameter 
by 14 inches long to represent each day’s work. The date upon 
which these are cast must be shown in raised figures on each test 
piece. The test pieces should be cast at different times during 
the pouring period, and the analysis of the test pieces must agree 
(within reasonable limits) with castings made that day. Castings 
not within the weight limits given on the drawings and failing 
to meet the above requirements will be rejected and returned. 

Cylinder Stresses and Wall Dimensions—In calculating 
the maximum stresses which may occur in a gasoline engine 
cylinder, account must be taken of the fact that if one or more 














76 


THE CYLINDER. 


charges should be 
“missed” (i. e., fail to ig¬ 
nite) all of the spent gases 
will be pumped out of the 
cylinder, and the next 
charge to fire will be 
stronger in the proportion 
of the piston displacement 
plus compression volume 
to the piston displacement. 
Therefore, although the 
normal explosion pressure 
in an automobile motor 
seldom exceeds 300 
pounds per square inch, it will be well to figure on abnormal ex¬ 
plosion pressures of 400 pounds per square inch. 

Now, consider a section of a cylinder of b inches bore 1 
inch long (Fig. 33). The pressure developed in the cylinder 
by the explosion tends to rupture the cylinder wall along 
lines parallel with the centre line and at opposite ends of a 
diameter. The rupturing force for the section of the cylinder 
considered is evidently 400 b pounds (considering the explo¬ 
sion pressure to be 400 pounds per square inch). If the cylinder 
wall has a thickness t and its material has a tensile strength 
of 24,000 pounds per square inch, the resistance to rupture of 
the two sections 1 inch long and t inch thick is 

2x^X24,000 = 48,000 t pounds. 

Calling the factor of safety f, we have 



_48,000 1 


f= 


400 b 


= 120 


£ 

b 


If we allow a factor of safety of 4, then 


4 = 120 — 
b 

and 

t = — inch. 

30 

This formula when applied to cylinders of small bore gives 
values for the cylinder wall thickness which, while large 
enough so far as withstanding the stresses of explosion is 
concerned, would be too small from the standpoint of shop 
production. If the water jacket is cast integral, as it usually 
is, the cylinder can be machined only on the inside, and the 
minimum thickness of the wall then depends upon the accu¬ 
racy with which the cores are set. Some allowance must be 
made for inaccurate core work. From the data of 18 modern 










THE CYLINDER. 


77 


motors of both American and European construction the fol¬ 
lowing formula for wall thickness has been derived: 

t = — + s inch .(32) 

While the motors from which the data were taken ranged 
in bore only from 3 to 5 inches inclusive, the formula may 
safely be applied to a much wider range of sizes. The nearest 
size in thirty-seconds of an inch to the result given by the 
formula should be chosen. If the cylinder becomes worn in the 
course of use, or damaged by lack of lubrication or on account 
of the piston pin coming adrift, it may be saved by regrinding. 
In the case of ordinary wear an increase in the diameter of 0.010 
inch is usually sufficient, but in cases of scoring the diameter 
must sometimes be enlarged as much as 0.040 inch. Some manu¬ 
facturers make it a practice to furnish over-sized pistons to be 
fitted into reground cylinders, and the Society of Automobile 
Engineers has standardized these pistons, which are made 0.010, 
0.020, 0.030 and 0.040 inch over-size. Wall thicknesses as given 
by the above formula admit of such regrinding. 

The cylinder head is usually made of the same thickness as 
the cylinder wall. If it were of hemispherical form its factor 
of safety would be the same as for the cylinder wall. The head 
usually approaches the hemispherical form near its edges 
where the strain is the greatest, and flattens out toward the 
centre. However, in T-head cylinders it is generally made 
flat, owing to the difficulty in bringing the clearance volume 
down to the required point. The water jacket usually extends 
down the cylinder to a point even with the top edge of the 
piston when in its lowest position. Some manufacturers, 
notably the German Daimler Company, make the jacket 
shorter, while a few make it longer. The width of the water 
space varies a great deal in different designs, and no relation 
between the width of the water space and the bore of the 
cylinder is recognizable in the data at hand. The widths vary 
between i 7 s and 24 inch. A good rule would be to make the 
jacket Yi inch wide for cylinders from 3 to 4 inches, 24 inch for 
cylinders above 4 and up to 5 inches and 24 inch above 5 inches. 
Liberal water spaces have the advantage that the core sand can 
be more effectively removed from the casting. 

The jacket wall is generally made as thin as the foundry process 
permits of. It can, of course, be made thinner in a small cylin¬ 
der than in a big one, because in the former the area is smaller. 
The average practice is to make the jacket wall thickness 



78 


THE CYLINDER. 


£2 inch for a 3 inch bore; 

& inch for a 4 inch bore; 

3*2 inch for a 5 inch bore; 

% inch for a 6 inch bore. 

On top of the cylinder a liberal water space is generally al¬ 
lowed, from 24 to 1^2 inches, according to the cylinder bore. 

In a single cylinder the cylinder wall is sometimes made 
inch thicker below the jacket, or at least it is gradually thickened 
as the flange at the lower end is approached. The lower end of 
the cylinder does not have to withstand the direct stresses of ex¬ 
plosion, but instead it is subjected to the tension due to the explo¬ 
sion pressure against the cylinder head and to the shearing and 1 
bending stresses due to the torque reaction or to the side thrust 
of the piston against the cylinder wall. It has been shown by 
James Angelus that if the cylinder wall is made sufficiently thick 
to withstand the direct bursting pressure of the explosion it is 
sufficiently strong to withstand the combined tension, bending and 
shearing stresses near its lower end; but, possibly owing to 
strains set up in the metal at the joint in cooling, and possibly 
because one or more of the nuts on the bolts may loosen, single 
cylinders sometimes break at the flange, and a slight increase in 
thickness toward the bottom end seems advisable. 

A single cylinder is secured to the crank case by means of four 
studs and nuts equally spaced. The proper diameter for the 
studs may be found by means of the equation 

d=b - + 1 s inch .( 33 ) 

Where the nuts come there should be bosses on the flange % to 
*4 inch high, finished off to insure a square seat for the nuts. 

Compression Space—We will suppose that the bore and 
stroke of the motor and the type of cylinder have been decided 
upon. The power output to be expected from a cylinder of 
given dimensions and the relation of bore to stroke will be dis¬ 
cussed in later chapters. A design for a T-head cylinder of 4 
inch bore and 5 inch stroke is shown in Fig. 34. The two valves 
are shown to have a clear diameter of i 24 inch each. We will 
suppose that the compression ratio is 4.25. The piston displace¬ 
ment volume is 

4x4x5x0.7854 = 62.83 cubic inches, 
and with a compression ratio of 4.25 the volume of the compres¬ 
sion space must be 

_ 62 : . ? 3 _ — ig.33 cubic inches. 

4.25 — 1 

In laying out the compression space care must, of course, be taken 




THE CYLINDER, 




iSeciJon at X~Y 


Fig. 34.— Single T-Head Cylinder. 









































































































80 


THE CYLINDER. 


that there is a free passage for the gas from and to the valves. 
With a clear diameter of the valve port of \Y\ inches, the outside 
diameter of the valve head would be, say, 2 x /s inches, and the 
valve stem would be j 7 a inch in diameter. The effective opening 
of the valve port would then be 

(i —tV) 0.7854 = 2.25 square inches. 

Now, one-half of the valve head is surrounded by the valve 
pocket wall at a fixed distance, and in order that the annular space 
between the valve head and the valve pocket wall may be equal 
in area to the corresponding portion of the valve port area, the 
diameter x of the valve pocket must be such that 

(x 2 — 2 '/s 2 ) 0.7854 = 2.25, 

from which we find 

^• = 2.71 inches. 

It is, however, not necessary to have the annular space quite 
as large as the valve port area, for as soon as the incoming gas, 
for instance, has passed through the valve port it is drawn toward 
the cylinder, and less passes around the outer half of the valve 
head than around the inner half. We will, therefore, make the 
valve pocket of 2^6 inch diameter instead of 2.71 inches. This 
will leave a free space of Ft i nc h all around the Valve head. 

The form of the compression space is shown in Fig. 35. The 
area of one head of the compression space is 27.81 square inches. 
If we made the space of a uniform height of Y\ inch, then the 
volume would be 20.86 cubic inches, or 1.53 cubic inches more 
than required for the compression volume. In order, therefore, 
to get the right compression volume the piston when in its top¬ 
most position should extend into this space a distance 

- 1 ‘ 53 . = y s inch approximately. 

12.56 

It is very desirable that the piston should over-travel the top 
•end of the finished bore a short distance, so that no shoulder 
may be worn in the bore. An over-travel of % inch will be 
about right. 

The cross section of the passage connecting the valve cham¬ 
ber with the cylinder is 

25/s x *4 = 1.97, 

or nearly 2 square inches, which is slightly less than the area of 
the valve seat port. For T-head cylinders it is good practice to 
make the inner diameter of the valve pocket 50 per cent, greater 
than the clear valve diameter and the height of the pocket 40 
per cent, of the clear valve diameter, or a little more, according 
to the compression ratio desired. 

While on this subject, it may be of interest to calculate the 



THE CYLINDER. 


81 



K ^ "l 00 

I- 3 !- ^ 1 - 1 

l 4 UJ. 35 - 

surface area of a compression chamber of this form, as com¬ 
pared with a compression chamber of cylindrical form of about 
the same diameter as the bore. As already stated, the head of 
the compression chamber has an area of 27.81 square inches. 
The circumference of the compression space is approximately 
22.35 inches, which with a height of inch gives a circumferen¬ 
tial surface of 16.76 square inches. Adding this to the 55.62 
square inches of the heads gives a total of 72.38 square inches 
area. On the other hand, if the combustion space was in the 
form of a cylinder of 4 inches diameter, the required height 
would be 1.56 inches, and the surface area would be 44.71 square 
inches. With the T-head the surface area of the combustion 
chamber is, therefore, about 62 per cent, greater than it should 
be with a valve in the head motor and cylindrical combustion 
chamber. 

Where a high compression ratio is to be used, together with 
large valves, it is sometimes difficult to get the combustion cham¬ 
ber sufficiently small. One of the expedients resorted to in such 
cases is to have the cylinder head slope downward toward the 
centre. In the above calculation of the compression space it was 
assumed that the openings over the valves are closed by plugs the 
lower faces of which come flush with the inner surface of the 
valve chamber wall. Quite frequently these plugs are chambered 
out on the inside, in which case proper allowance must be made 
for these chambers in calculating the compression space. 

Length of Cylinder Bore—The top end of the piston must 
pass slightly beyond the top end of the cylinder bore at the end 
of the upward stroke, for the reason already explained, and the 
























82 


THE CYLINDER. 


bottom end of the piston should also pass slightly out of the 
cylinder bore at the end of the downward stroke, partly to pre¬ 
vent the wearing of a shoulder at the bottom end and partly be¬ 
cause, if splash lubrication is employed, lubrication of the cylin¬ 
der wall will be facilitated. The over-travel at the bottom end 
may be made somewhat greater than the over-travel at the top 
end. Since we decided on l /$ inch for the latter, we may choose 
inch for the former, making the total over-travel inch. 

The length of the finished portion of the cylinder bore is evi¬ 
dently equal to the length of the stroke plus the length of the 
piston minus the total over-travel. In our example, if we assume 
that the piston is 4j4 inches long, the total length of the bore 
will be 

5 + 4^ — V2-9 inches. 

At the lower end the cylinder is always counterbored to a taper, 
as shown in Fig. 34. The object of this is to facilitate the intro¬ 
duction of the piston rings and of the piston itself. The greatest 
diameter of the counterbore should be slightly larger than the 
outside diameter of the piston ring in its free state. When cyl¬ 
inders are cast singly they are often turned with a cylindrical 
projection or guide at their lower end below the base flange, as 
indicated by dotted lines in Fig. 34. This base flange permits 
of accurately centring them in the crank case, which is bored to 
a corresponding diameter. If no such guide is provided the cylin¬ 
der should be held in alignment by means of dowel pins. 

The dimensions of the various parts, as calculated by means of 
the preceding formula, are shown in Fig. 34. The water jacket 
head is shown domed, which makes the cylinder neater in ap¬ 
pearance, and also adds to the strength of the head. Two bosses 
are shown cast on the jacket wall near its lower end. The open¬ 
ings are used in the first place for core prints for supporting the 
water jacket core. One of the openings can be used later for the 
cooling water inlet, and the other can be plugged. Instead of 
having the cooling water inlet at the bottom of the jacket it may 
be located underneath the exhaust valve chamber, and many de¬ 
signers consider this the preferable location, as it insures the 
most intense cooling effect near the exhaust valve, where ex¬ 
cessive heating is likely to do damage first. However, in case 
the water enters the jacket near the exhaust valve, a drain cock 
must be placed at the bottom of the jacket space, so that all water 
can be drained out of the jacket when desired. 

*In the design Fig. 34 the valves are entirely surrounded by 
liberal water jackets, which is now considered an essential point. 
Where the valve chamber wall approaches closest to the cylinder 


THE CYLINDER. 


83 


wall the cylinder water jacket space is necessarily somewhat re¬ 
duced. The width of the water space at this point should not be 
less than 54 inch, as otherwise there may be difficulty in the 
foundry. In laying out the water spaces the designer must see to 
it that there is no point from which the water cannot reach the 
outlet without passing to a lower level, as otherwise steam pockets 
will be formed, and the particular portion of the cylinder wall 
thus deprived of water may overheat. 

Twin Cylinders—In the design of twin cylinders, about the 
first question that comes up is whether it is necessary to have 
a water space between the cylinders at the point where they 
are nearest to each other or whether a common wall may be 
used at this point. Both arrangements are practicable, and both 
are in extensive use, although in this coun¬ 
try by far the greater number of twin 
cylinders have a water space between the 
cylinders. In Europe, on the other hand, 
many engines are made with a common 
cylinder wall. If the latter style of con¬ 
struction is adopted it is necessary to pro¬ 
vide an extra thickness of metal where the 
wall is common to both cylinders. By re¬ 
ferring to Fig. 36 it will be readily seen 
that the thinner this wall is made the 
greater will be the distance from the Fig. 36. 

centre of it to the water jacket; and, be¬ 
sides, the cross section of the metal through which the heat 
must travel will be less, while the area through which heat 
is absorbed will be greater. A good rule is to make the 
minimum thickness of the common wall 50 per cent, greater 
than the thickness of the regular cylinder wall.. Some manufac¬ 
turers make it twice as thick as the regular cylinder wall. Among 
the makers using a common wall are Wolseley in England, Adler 
in Germany and Midland in this country. 

A typical twin cylinder design is shown in part sectional plan 
and elevation in Fig. 37. These cylinders are also of the T-head 
type. That part of the cylinder wall which is exposed to the ac¬ 
tion of the hot gases is completely surrounded with water for 
the greater part of its length, but the two compression chambers 
have a common wall. One respect in which this design differs 
from that of Fig. 34 is that* the valve passages underneath the 
valve seats, instead of turning through a right angle, turn through 
an angle of only 45 0 . At the outlet of the valve passages there 
are oval bosses A for pipe connections by means of flanged fit- 




84 


THE CYLINDER. 


tings. The 45 0 angle gives a freer passage for the gases, and is, 
therefore, advantageous. The jacket walls in this design are 
provided with large hand holes B at opposite ends. This permits 
of a more substantial support for the jacket core in casting, and 
also greatly facilitates the removal of the core sand. It is quite 
important that all of the core sand be removed from the casting, 
as otherwise it is likely to become loosened later, when the motor 
is in service, and choke the water passages. The two hand holes 
are closed by suitable cover plates. An increased thickness of 
metal is provided all around the hand hole and the flange thus 
formed is milled off to insure a water-tight joint with the cover 
plate, which is held in place by means of six screws. 

In this design the two valves oh each side are siamesed, that is, 
the spaces underneath the valve seats communicate with each 
other, and there is only a single outlet for the two valve cham¬ 
bers and only a single connection. There is some objection 
to siamesing the exhaust valves in this way, in the case of 
two and four cylinder motors, because in these motors one cylin¬ 
der begins to exhaust before the exhaust valve of the preceding 
cylinder closes, and if the two exhaust valves are siamesed it is 
inevitable that some of the exhaust gas from one cylinder will 
blow over into the other, thus preventing its complete evacuation. 
It is, however, common practice to Siamese the exhaust valves also. 
In six cylinder motors with twin cast cylinders the firing order 
can be so arranged that the two cylinders of any one pair will 
never fire in direct succession. The exhaust opening periods of 
the two cylinders which are cast together will then be separated 
by 240°, which is more than the period of opening of one ex¬ 
haust valve. 

In this design the water enters the jacket space through the 
opening C at the side of the exhaust valve chamber just above 
the exhaust outlet, and it leaves the jacket on top midway be¬ 
tween the two cylinders where a boss D for a pipe connection 
is provided. At the sides of the cylinders a short distance under¬ 
neath the bottom of the water jacket are provided little projec¬ 
tions E which serve as supports for a tool used for compressing 
the valve springs when it is desired to remove the valves. These 
valve tool supports are quite a convenience and should be more 
generally provided. Between the two valve chambers on each 
side is shown a boss F, which is drilled for the reception of a 
stud. This stud serves to clamp in place a yoke by means of 
which the plugs or caps over the two valves are held down on 
their seats. 

A twin cylinder is always secured to the crank case by means 
of six studs or bolts which are located as shown in Fig. 37 * The 


THE CYLINDER 


85 



Fig. 37 — Twin T-Head Cylinder, 




























































































80 


THE CYLINDER. 


diameter of these studs can be calculated by means of the for¬ 
mula given for single cylinders. The two small plugs G serve 
to close holes left by core prints for the jacket core. 

L-Head Twin Cylinder —An L-head twin cylinder design is 
shown in Fig. 38. The two inner valves in an L-head motor are 
practically always the inlet valves, because, as already stated, 
siamesing is more advantageous in the case of the inlet valves 
than in the case of exhaust valves, and in an L-head motor only 
the inner pair of valves can be siamesed. Besides this, placing 
the two inlet valves together lessens the need for a water jacket 
between the two valve chambers, and in an L-head motor the 
space thus gained can be used to advantage for increasing the 
valve diameters. 

The valves are placed as close together as possible. If the 
clear diameter is d the distance apart of the valve centres will 
be about 1.44 d, if the valve openings are closed by separate 
plugs or caps and the water jacket is brought quite close to the 
valve seats where they approach closest to each other. In the 
design here shown the clear diameter of the valves is 0.4 b, and 
it will be noticed that the valve box is of the same length as the 
cylinder block. If larger valves are desired the valve box will 
be longer than the cylinders, and designs of this kind are not 
uncommon. An expedient resorted to in order to bring the 
water as close to the valve seat as possible at the point where 
the two valve pockets have a common wall, without sacrificing 
valuable space in a longitudinal direction, consists in making the 
inlet pocket rather shallow, turning the inlet passage sideways 
directly underneath the valve seat. 

The outlets of the valve passages are all on the same level, 
and bosses A are provided in the casting, to be drilled and tapped 
for the reception of studs by means of which the inlet and ex¬ 
haust manifolds are clamped in place. One feature of this design 
is that the water space around the cylinders is made larger at the 
top end of the cylinders than at the bottom. This insures a 
greater amount of water where the most heat is absorbed, which 
is considered advantageous by some designers. It also gives a 
rather neat form of cylinder. Both the cylinder head and the 
jacket head are cast integral in this design, and there is a 
central hole in the head over each cylinder into which can 
be screwed the compression relief cock. There is a large hand 
hole in the side of the jacket, which is shown closed by a 
sheet metal plate held in place by fillister head screws, water 
tightness being insured by a rubber gasket. This provides 
for the support of the jacket core, and two additional core prints 


THE CYLINDER. 


87 




v Section on V—W 


Fig. 38 —Twin L-Head Cylinder. 
















































































































88 THE CYLINDER. 

are provided for at the ends of the valve box, as indicated by 
plugs at these points. 

Another feature in cylinder construction illustrated in this cut 
is an oil groove C cut in the cylinder wall about even with the 
centre of the piston pin when the piston is in its lowest position 
Ordinarily oil grooves are cut into the wall of the piston, but 
some designers contend that the proper place for them is the 
cylinder wall. This oil groove is uncovered by the piston when 
the latter approaches the upper end of its stroke, and the groove 
will then be filled with oil from the splash, which will be dis¬ 
tributed over the piston as the latter makes its downward stroke. 
The oil groove has a tendency to concentrate the strains, and in 
order to prevent any weakening of the cylinder it is advisable to 
reinforce the wall at that point, making it of 25 per cent, more 
than the normal thickness. The lugs for the cylinder bolts are 
shown strengthened by the ribs D, a practice followed by some 
French designers. This design is hardly adapted for American 
conditions, however, owing to the fact that considerable hand 
chipping may be necessary to get a good square seat for the 
washers. Heavy washers or spacers must be placed under the 
nuts to bring the latter sufficiently high to be conveniently 
reached with a wrench. In large bore cylinders with thin walls, 
four vertical ribs are sometimes provided on tb#> cybnder wall 
from the lower end of the jacket to the base flange. 

Separate Cylinder Heads.—Cylinders with valves in the head 
necessitate the use of valve cages, except when, as in tne case 
of the Knox motor, the cylinder head is removable. In the 
early automobile motors the cylinder heads were invariably 
cast separate from the cylinders and bolted to them, with a 
gasket of asbestos cloth between. Segmental openings were 
cut into the gaskets, through which the cylinder water space 
communicated with the head water space. But as the wall 
thicknesses were made smaller and smaller, it became more 
difficult to keep these gaskets in place by the pressure be¬ 
tween the cylinder and head. The clamping surfaces were 
faced off in the lathe with a coarse feed and the gasket ma¬ 
terial was rendered more tenacious by interweaving it with 
copper wire; nevertheless the gaskets would frequently blow 
out under the pressure of explosion or burn out when the 
water circulation failed, and renewing them was one of the 
most dreaded repair jobs in those days. This led to the prac¬ 
tical abandonment of separately cast heads. They were rein¬ 
troduced some years ago by the Knox Automobile Company 
because they permit of forming valve seats directly in the 
head, doing away with cages, thus permitting of using larger 


THE CYLINDER. 


89 



Fig. 39.—Knox Cylinder Head. 

(Water connections swung around 90 degrees into the plane of the valves.) 

valves, insuring more effective cooling of the valve seats and 
rendering the valve construction simpler. Gasket difficulty 
is obviated by placing the cylinder and head jacket spaces in 
communication through an outside, U-shaped pipe fitting. 

Recently quite a number of four cylinder block motors have 
been made with separate heads, an asbestos filled sheet copper 
gasket being used between head and cylinder block, which latter 
is often cast integral with the top half of the crank case. These 
motors do not always have the valves in the head, the chief object 
in making the head separate being to facilitate the operation of 
decarbonizing the cylinders. 

Valve=in=Head Cylinders—The ordinary type of valve-in¬ 
head motor with integral cylinder heads embodies cages in 
its construction. These consist of cylindrical shells which fit 
into corresponding bored chambers in the cylinder head, being 



Fig. 40.—Jackson Cylinder Head. 







































































yo 


THE CYLINDER. 


forced against a square or inclined shoulder near the bottom 
of the valve cage pocket by means of an annular nut screwed 
into the outer, threaded end of the pocket, or by means of 
studs passing through a flange on that portion of the valve cage 
outside the bore, and into the cylinder head wall. The Jack- 
son cylinder, Fig. 40, is typical of this construction. Within 
the valve cage there is a spider which supports the valve 
guide and in the wall of the cage there is a circular or square 
opening by means of which the interior of the valve cage 
communicates with the valve passage cast in the cylinder 
head. The cage at its inner end is usually ground to a coni¬ 
cal seat and a gasket is placed between the cage and the nut. 
In multicylinder., valve-in-the-head motors, in which the valves 
are arranged vertically, they are usually placed all in one 
row; that is, the two valves in each cylinder head are placed 
side by side in a fore-and-aft direction. But where the valves 
are inclined they are placed crosswise of the cylinder. 

Sheet Metal Water Jackets—Water jackets of sheet metal 
have two advantages. They are lighter than integrally cast 
jackets, because they can be made thinner; and, besides, they 
cannot easily be injured by freezing of the jacket water. 
These sheet metal jackets are generally spun or pressed of 
copper. The two most prominent users of such jackets have 
been the Cadillac Automobile Company and the Chadwick 
Engineering Works. The former cast its cylinders singly, 
while the latter cast them in pairs. Panhard & Levassor, 
of France, have also used copper jackets for some of 
their motors. The Cadillac cylinder is here illustrated as an 
example of this form of construction. The cylinder is cast 
with a radial flange A at the bottom of the jacket space. This 
flange is slightly tapered and is turned off very coarsely, and 
the copper jacket is pressed over this. A steel ring about two 
one-thousandths inch small is then forced over the flange and 
insures a tight joint and one which enables the jacket to be 
quickly removed. The valve pocket is formed integral with 
the cylinder head, and the head is held to the cylinder by means 
of a right and left screw nipple. The under side of the head 

is provided with two V-shaped circular ridges which are 
forced down at B and B on each side of the water jacket ports, 

thus making a tight joint. 

If a sheet nietal jacket is to be used, account must be taken 
of the difference in the expansion of the cylinder and the 
jacket when the motor is started. When the jacket is in the 
form of a cylinder with flat or domed head, held at its open 


THE CYLINDER. 


91 


end and at the centre of the head, it is free to expand and 
contract longitudinally, but when it is in the form of a cylin¬ 
drical shell, held at both ends, it must be provided with cir¬ 
cular corrugations, or trouble will be experienced from leaky 
joints. 



Fig, 41.—Cadillac Cylinder, With Sheet Copper Jacket. 


Block Cylinders—About 1906 the practice of casting four 
cylinders in a single block originated in France, and though 
at first it was predicted that the number of “wasters” (imper¬ 
fect castings) would be so large that this practice would be aban¬ 
doned again, it has constantly gained ground. 

The simplest form of block cyclinder is that in which the cylin- 































































92 


THE CYLINDER 





Fig. 42.—Block Cylinder Design, With Integral Exhaust Manifold. 
















































































































THE CYLINDER. 


93 



Fig. 42-B. —Section on M-N. 



Fig. 42-A. —Section on X-Y. 






























































































94 


THE CYLINDER. 


der head is cast separate and no manifolds are incorporated in 
the cylinder casting. Slightly more complicated is the cylinder 
block with integral head, but the most intricate piece of foundry 
work is a four or six cylinder motor with integral manifolds. 
The latter type of cylinder casting results, howev.er, in the neatest 
and most compact motor, and is strongly favored for small 
motors. 

Fig. 42 shows a block cylinder design in which the exhaust 
manifold is cast integral. The valves are all on one side and the 
inlet ports are siamesed as usual. The pockets directly under¬ 
neath the exhaust valves connect by straight horizontal passages 
with the longitudinal exhaust passage. As clearly shown in the 
section X-Y, this exhaust manifold is surrounded by cooling water 
on three sides. Where the exhaust manifold is cast integral, a 
relatively larger amount of heat goes into the cooling water, and 
in view of this fact exceptionally large water spaces are provided 
around the valve passages in this design. The valve pockets are 
entirely surrounded with water. The water enters the jacket 
space underneath the longitudinal exhaust passage at the middle 
of the length of the motor, and leaves through the jacket top 
plate, which is developed in the form of a water outlet con¬ 
nection. 

Another form of block cylinder is shown in Fig. 43. In this 
case both the exhaust and inlet manifolds are cast integral with 
the cylinder. The longitudinal exhaust passage is less completely 
water-jacketed than in Fig. 42, and is located outside the valve 
box proper. Cooling flanges extend along both its top and outer 
side wall. The longitudinal passage forming the inlet manifold 
runs along the lower outward edge of the valve box. With a 
design of this kind a stronger water circulation would be re¬ 
quired than with that of Fig. 42. This illustration also shows a 
cover over the valve springs, which it is usual to provide in 
motors of this type. 

Abroad it is the prevailing practice to locate the carburetor 

on the right hand side of the motor in automobiles. In the case 
of T-head motors the inlet valves are practically always on that 
side, and in the case of L-head motors, with all valves on the 
left hand side, the inlet pipe often runs across the top of the 
cylinders or through between the two pairs below the water 
jacket. With block motors it is naturally desired to avoid such 
long intake pipes, which look ungainly and are more or less im¬ 
practicable on account of the condensation of gasoline taking 
place in them in cold weather. It therefore occurred to a num¬ 
ber of designers to cast an inlet passage integral with the cylin¬ 
der block extending from the left to the right hand side. In one 


THE CYLINDER. 


95 



Fig. 43.—Section of Block Cylinder With Integral 
Inlet and Exhaust Manifolds. 


or two motors this passage is carried around the end cylinder, 
but the more usual plan is to make the distance between the 
two middle cylinders a little larger than ordinarily and core an 
inlet passage between them. This is done in the new Fiat “35,” 
and also in the Sampson “30” motor, Fig. 44. Since the wide¬ 
spread adoption of left hand drive, many L-head motors are cast 
with the valve pockets on the right hand side. Some designers 
then place the carburetor on the right hand side, while others 
carry the inlet passage or pipe to the left side of the motor. 

Pattern Making and Molding—In the design of the cyl¬ 
inder it is well to consult with the pattern maker, because a 
cylinder is at best a difficult piece to mold, and the advice of an 
experienced mechanic may obviate trouble later. Cylinder pat¬ 
tern making has become an art by itself, and it is a good plan 
to trust a job of this kind only to a man experienced in this 
particular line. 

Cylinders must always be molded with the head downward, 
for the reason that blow holes, porous spots, etc., are most likely 
to occur near the top of the casting, and the head of the cylinder, 
which is the working end, must of necessity be of sound metal. 





















































H^ The Hor»eln» Age 1 -—- 

Fig. 44.—Sampson Quadruple Block Cylinder Casting. 














































































































































































THE CYLINDER. 


97 


In order that the metal may flow freely to every part of the mold 
the latter must be thoroughly vented, and a high riser (opening 
on top which fills with metal after the mold is filled) is provided, 
so that the molten metal in the mold may be subjected to the 
pressure due to a considerable head, and all gases forced out 
through the vents. It is imprisoned gases that cause blow holes 
and other imperfections. 

It is not often that the compression space is of such a simple 
geometrical form as in the cylinder Fig. 34, and when it is quite 
irregular it is naturally more difficult to accurately calculate the 
compression volume provided by the design. It is then a good 
plan to let the pattern maker make a wooden model of the com¬ 
pression chamber. This can then be weighed and its weight 
compared with that of a simple geometrical form (cube or 
prism) made of the same kind of wood, which enables its vol¬ 
ume to be obtained by the solution of a simple geometrical 
proportion. 

Testing Castings—When the castings have cooled they are 
cleaned of core sand, the seams, etc., are chipped off. and the 
castings are then pickled or cleaned of scale in a sulphuric acid 
solution. 

When the castings are received at the factory they are care¬ 
fully inspected and tested for imperfections before any machine 
work is done upon them, so that no labor may be wasted on im¬ 
perfect castings. The usual test is by water pressure, and a fixture 
and stand for making such a hydrostatic test rapidly has been 
described by Curtiss Smith. The fixture, Fig. 45, consists of a 
skeleton frame designed to receive the cylinder casting. The 
frame is provided with two trunnions by means of which it is 
supported on the standard, Fig. 46, and which permit of tipping 
it to any angle, for examining the bore and all parts of the 
casting. The castings are held in place in the frame by means 
of three thumb screws, two pressing laterally against the lower 
unjacketed portions of the cylinders and the third centrally 
against the valve chamber. All holes into the jacket must be 
plugged, except the one through which the hydrostatic pressure 
is applied. The design of the fixture would naturally vary 
with the form of the casting to be tested. The one here shown 
was designed for a twin cylinder casting with central openings 
in the cylinder head and jacket head. The hole in the cylinder 
head is plugged by a soft rubber washer £ on a stem F, which 
is held in place by a copper washer G and cotter pin H. This 
washer is backed up by a steel thrust disc 7 , which has a copper 
washer / interposed between it and a shoulder on the spindle. 


98 


THE CYLINDER. 



Fig. 45.—Cylinder Testing Fixture. 


The stem F is threaded at its lower end for applying pressure 
and is squared at its upper end for the operating handle. 

The hole in the jacket head is stopped by a soft rubber plug 
K held in a thrust block L with retaining flange M and copper 

washer N. This plug is operated by the hollow spindle O, 

• 

which screws through the cast frame and carries the spindle F. 
The spindle O extends beyond the handle P to form a stuffing 
box, being threaded at its upper end to receive the gland nut Q. 

On the back of the cylinder casting for which this fixture 
was designed there is a large square hand hole. For plugging 
this hole and providing a water connection a special construc¬ 
tion is used. The trunnion at this side of the fixture is made 
of steel, bolted to the casting and drilled and tapped to receive 
a clamp screw. Grooves Y are planed between the trunnion 
lugs, in which the block T slides. This block is faced with 
a soft rubber packing, which is held in position by flathead 
screws. The block T is bored and tapped to receive a pipe 
through which the testing water enters the jacket space. Con¬ 
nection to the water supply is made by means of a union with 
a short section of pipe and rubber hose. 

The standard in which the above fixture is mounted consists 
of two end castings AA held in place by spacers B. The trun¬ 
nion bearings are provided with clamping, caps C which are 





































































































THE CYLINDER. 


99 


hinged at D and clamped with the T handle E. The standard 
is made high enough to allow all parts of the testing fixture ex¬ 
cept the plug stems to pass over the spacers B. 

Boring and Grinding Cylinders—The first machining opera¬ 
tion on a cylinder casting consists generally in milling off the 
base and head flanges and the bosses for the different pipe con¬ 
nections. The base flange is milled first because the cylinder is 
usually secured to the boring fixture by means of this flange, 
at least in the case of twin and block cylinders. The next op¬ 
eration is the rough boring. This is done in either a horizontal 
or vertical boring mill. Generally two cuts are taken to re¬ 
move the stock allowed for finishing, but the cylinder bore is 
left about o.oio inch small in diameter, this amount of stock 
being removed by grinding. 

The method of holding the cylinders to the carriage of the 
boring mill differs considerably in different shops, and also with 
the type of cylinder casting to be bored. Single cylinders are 
often held in a cylindrical fixture, the upper half of which is 
hinged to the lower, so that the castings can be easily put into 
the mill and removed. In the case of twin and quadruple cylin¬ 
ders the general practice is to bolt the casting by means of its 
base flange to an angle bracket which is bolted to the lathe car¬ 
riage. A number of special machine tools have been designed 
for boring automobile cylinders. Some of these have revolv¬ 
ing bed plates of double ended form, so that a casting can be 
clamped to one end while another is being bored, and the ma¬ 
chine thus be kept in operation almost continuously. A typical 
cylinder boring machine is shown in Fig. 47. 

In boring a cylinder which is open at both ends a boring bar is 



Fig. 46.—Cylinder Testing Standard. 







































100 


THE CYLINDER 





Fig. 47. —Typical Cylinder Boring and Milling Machine (Beaman & Smith). 













THE CYLINDER. 


101 


used which is supported between centres in a lathe or boring ma¬ 
chine, and is provided with a series of cutters which are equally 
spaced about the circumference and along a portion of its length. 
The boring bar is only little smaller in diameter than the rough 
cylinder bore, and is provided with a mechanism for moving the 
cutters outward. At present, however, practically all cylinders 
have integral heads, and these must be bored by means of a 
spindle supported at one end only. It is quite a problem to give 
a sufficiently rigid support to this spindle. Fig. 48 shows a cylin¬ 
der boring machine spindle patented by the Beaman & Smith 
Company, of Providence, R. I. This spindle is designed for use 
on a vertical boring machine. A is the hollow spindle which is 
driven through the gear wheel D, secured to its upper end, and 
has the boring and chamfering tools W, X, Y screwed to its lower 
end and secured by a set screw. The spindle A turns on the 
stationary stud C, being provided with bronze bushings B B. The 
end thrust due to the feed is taken up on ball thrust bearings 
E, G. The chamfering tools are held in a split ring X, which is 



Fig. 48— Beaman & Smith Boring' Machine Spindle. 























































































102 


THE CYLINDER. 


clamped to the spindle A. The spindle, gear, housing, etc., are 
carried on a long vertical saddle with adjustable stops. The firm 
support of the spindle is apparent from the drawing. 




Fig. 49. —Expanding Tool for Finishing Compression 

Chamber. 

An arrangement for counterboring the compression chamber is 
illustrated in Fig. 49. This consists essentially of a spindle A 
with six inserted cutters B, with which the finishing cut is taken. 
The spindle is split longitudinally into six equal sectors over more 
than half its length. When the top edge of the cutters B reaches 




































THE CYLINDER. 


103 


the bottom end of the regular bored portion of the cylinder the 
spindle is expanded by the conical frustum C, which is secured 
to the fixture by means of which the cylinder casting is held to 
the bed plate of the boring machine, and in this way the head 
end of the cylinder can be turned several hundredths of an inch 
larger in diameter than the regular bore. At the same time as 
the combustion chamber the piston ring chamfer is turned by 
means of the tools D. 

In the early years of the industry it was customary to finish 
cylinder bores in the lathe, four or five successive cuts being 
taken, the last with a very broad nosed tool, which gave a com¬ 
paratively smooth finish. The trouble with this method is that 
the cast iron of the cylinder is not of uniform hardness. There 
are hard spots and soft spots in the iron, and since the wall of 
the cylinder is very thin, when the tool encounters a hard spot 
either the cylinder wall or the tool will spring slightly, and a 
high spot result. Now, in an engine cylinder, what is required 
is an absolutely cylindrical bore, which is the only kind that 
will insure a gas-tight fit of the piston rings. To improve upon 
the original method the lapping process was introduced, which 
consists in working a dummy piston back and forth in the bore 
of the cylinder with some abrading material, like ground glass, 
mixed with oil and smeared over the cylinder wall. The ob¬ 
jection to this method is that there is danger of some of the 
abrasive being forced into the metal of the cylinder wall, and 
starting cutting of the wall when the cylinder is in service. The 
smaller sizes of cylinders are frequently reamed after the final 
cut has been taken in the lathe. The latest process for finishing 
cylinder bores, and which is now coming into very extensive 
use, consists in internal grinding. 

After the roughing cuts have been taken the cylinders are 
frequently annealed. This relieves the internal strains due to 
shrinkage, etc., and renders the metal of more nearly uniform 
hardness. If the castings are not annealed they are generally 
allowed to age or season for a period varying from a few days 
to several weeks. The object of this “seasoning” is practically 
the same as that of annealing. 

Cylinder grinding machines have been developed by a num¬ 
ber of firms, including the Heald Machine Co., of Worcester, 
Mass., and the Brown & Sharpe Mfg. Co., of Providence, R. I. 
In these grinders the twin or multiple cylinders are clamped 
to a carriage similar to a lathe carriage which has a longitudinal 
feed and a cross adjustment. The grinding wheel and spindle 
are carried in a rotating head which is provided with a com- 


104 


THE CYLINDER. 



Fig. 50— Brown & Sharpe Cylinder Grinder. 

bination of eccentrics, by means of which the wheel can be fed 
against the work or set for bores of different diameter. These 
grinders are provided with several speed changes, a highor speed 
being usually employed for the finishing cuts. In the case of a 
twin or multiple cylinder, one bore is finished first and the 
casting is then moved sideways by means of the movable carriage 
to bring the second bore in line with the head of the grinder, 
which can be accurately done by means of a micrometer adjust¬ 
ment. 

In ordinary grinding work the metal ground off and the parti¬ 
cles of abrasive lost by the grinding wheel are removed by a 
stream of water which is constantly played upon the parts being 
ground. In cylinder grinding an air suction is substituted for this 
stream of water, which carries the metal dust and abrasive away 
as fast as they are formed. In some shops water at a temperature 
of about 180 0 F. is kept circulating through the jackets while 
the grinding is being done, the object being to have the cylinder 
walls at as nearly working conditions as possible. It is claimed 
that if the cylinders are ground cold and an absolutely perfect 
cylinder is obtained, the bore will not be perfectly true when the 
cylinder is heated up. One disadvantage of grinding hot is that 
the bore will not be true when cold, and a perfectly ground cylin¬ 
der may be rejected by the inspector. It is best, therefore, when 
















THE CYLINDER. 


105 


grinding hot to inspect the cylinders while they are still hot. Cyl¬ 
inder bores are generally tested for accuracy by means of limit 
gauges. These consist of two steel plugs or cylinders dififering in 
diameter by one or two one-thousandths of an inch. The smaller 
one of the two should just enter the cylinders, and the larger one 
should not. 

Valve Plugs—The openings over the valves are usually closed 
by screw plugs which are chambered out to reduce their weight. 
These openings are bored to a diameter -fa or y$ inch larger 
than the outside diameter of the valve head, and are then 
threaded with a tap which usually has 12 threads to the inch. 
An asbestos filled copper gasket is placed underneath the flange 



Fig. 51.—Threaded 
Valve Plug. 



Fig. 52.—Plugs Held Down by Yoke. 




Fig. 53.—Winton 
Breech-Block 
Valve Closure. 


of the plug to insure gas tightness. Quite frequently the spark 
plugs are screwed into these valve plugs, and the plugs must 
then be drilled and tapped for the purpose. The most commonly 
used thread for spark plugs is the S. A. E. standard, which 
is % inch in diameter and cut with 18 threads to the inch. The 
part of the plug below the flange is 54 inch long. The priming 
cocks also are sometimes screwed into the valve plugs, but the 
preferred practice seems to be to place them in the cylinder head, 
so that the kerosene, etc., introduced through these cocks will 
be sure to get into the cylinder proper. 

When there is sufficient room between the valve plugs, as there 
generally is in T-head cylinders, a stud and yoke are often em- 















































106 


THE CYLINDER. 


ployed for holding two adjacent valve plugs in place. The neces¬ 
sary size of stud may be found as follows: Assume a maximum 
explosion pressure of 400 pounds per square inch and let the 
diameter of the valve plug be d. Then the total pressure 

on the plug is 400 v d * . Owing to the leverages, the tension on 

4 

the stud is twice this, and allowing a factor of safety of 5, and 
considering the tensile strength of the material to be 100,000 
pounds per square inch, the size of the stud must be such that 
the cross section at the bottom of the thread is 

2 X 5 X 400 X tt d* _rf* 

4 X 100,000 32 

For a plug 215 inches in diameter the area should be 

2A X 2rj . t 

_!« -Li = 0.15 square inch. 

32 

The nearest size to this in sixteenths of an inch, cut with an 
S. A. E. standard thread, is Vi inch. The yoke must be de¬ 
signed to have the same strength. It is preferably made of a 
drop forging. If no plugs or cocks screw into the valve plugs, 
the ends of the yoke can bear on a central boss on the valve plug, 
otherwise they must be forked and bear on the flange of the 
plug, leaving the centre free. The advantage of quick removal 
of the plugs with this method of fastening is lost if a spark 
plug is screwed into the valve plug. With no spark plug, 
all that is necessary in order to remove the valve plug is to give 
the nut on the stud a turn of two, swing the yoke around and 
lift the plug out. Such a plug can be ground to its seat and no 
gasket is then required. 

Another form of quick removable valve plug embodies the 
principle of the breech block. A disc or plate fits against a 
shoulder at the lower end of the valve opening, and a yoke with 
flattened ends is adapted to engage with its ends in opposite 
grooves milled in the wall of the opening, which it enters through 
a slot provided for the purpose. A set screw extends through 
the centre of the yoke, and by drawing this up tight against the 
disc at the bottom of the valve plug opening the disc is fixed in 
place. 

Cylinder Head Closures—It is comparatively seldom that cyl¬ 
inder heads are cast entirely closed, although if they are made 
that way the amount of machine work necessary is reduced. The 
reason for casting them with an opening is that the cylinder 
core can then be better supported and the wall thickness will 
be more uniform. The jacket heads are generally cast with 





THE CYLINDER. 


107 


large openings, especially in block 
cylinders, which are closed by alumi¬ 
num plates. Fig. 54 shows one rather 
popular method of closing the two 
openings in the top of the cylinder 
casting. A plug is screwed into the 
opening in the cylinder head. This 
plug is provided with an integral 
stud, by means of which the alumi¬ 
num cover plate is clamped down. 
Instead of making the stud integral 
with the plug it can be screwed into 



Fig. 54.— Cylinder Head 
Closure. 


it. The stud may be drilled through its centre for the reception 
of the priming cock. In some motors the openings in the cylinder 
head and jacket head are closed by separate plugs. 

Fan Bracket Support—In a considerable number of cars the 
bracket for supporting the radiator fan 
is fixed to the engine cylinder, and if 
this arrangement is desired provisions 
must be made in the cylinder design 
for the necessary screw holes and ma¬ 
chined surface. The bracket usually has 
quite a long overhang, and a rigid sup¬ 
port must therefore be provided, with 
a relatively long vertical base. Fig. 

55 shows a typical design. The bracket 
is held in place by substantial studs 
screwed into the cylinder wall near the 
top and bottom of the water space. 

The bracket itself is provided with ob¬ 
long holes through which these studs 
pass and with a set screw on top, the 
point of which bears against the top 
stud. By means of this set screw the 
bracket can be adjusted up and down to get the proper belt tension. 



Fig. 55.—Fan Bracket 
Attachment. 





































CHAPTER VII. 


PISTON, PISTON RINGS AND PISTON PIN. 

The piston in a gasoline motor serves a triple function. It 
forms the movable wall of the combustion chamber, allowing 
that chamber to increase and decrease in volume. It receives 
the pressure of the expanding gases and transmits this pressure to 
the connecting rod, and it forms a crosshead through which the 
side thrust due to the angularity of the connecting rod is trans¬ 
mitted to the cylinder wall. 

The piston makes of the combustion chamber a closed vessel 
of variable capacity. Since the pressure within the combustion 
chambe! s sometimes as high as 400 pounds per square inch, a 
high degree of gas tightness is essential. The cylinder, as al¬ 
ready stated, is bored so as to form as nearly as possible a true 
cylinder. If the piston were turned to exactly the same diameter 
as the bore over its whole length it might give a fairly gas tight 
joint. But this is impossible, owing to the difference in the heat 
expansion. The piston head heats more than the cylinder wall, 
because it is not water jacketed, and the head end of the piston 
also heats more than the open end, which latter is not exposed to 
the heat of the burning gases, and receives heat only by trans¬ 
mission. These facts must be taken into consideration in design¬ 
ing the piston, which must be made of less diameter at its head 
end than at its open end, and of less diameter than the bore at 
its open end. If it were made of the same diameter as the bore, 
so it was a tight running fit in the cylinder when cold, it would 
stick when the motor heated up. Since the piston must be made 
of smaller diameter than the cylinder it cannot be depended upon 
to insure gas tightness. This quality is attainable only by means 
of flexible, split metallic rings, called piston rings, three or four 
of which are placed upon the piston in grooves turned for the 
purpose in its outer circumference. These rings are originally 
turned to a somewhat larger diameter than the bore of the cylin¬ 
der, and a small section is then cut out of them, so that they can 


108 



PISTON, PISTON RINGS AND PISTON PIN. 109 


be sprung together to almost go inside the cylinder. Their outer 
surface is then finished off again while in this compressed state 
to the exact diameter of the bore. Since the ring was com¬ 
pressed to the diameter of the bore, when in its groove in the 
piston and inserted into the cylinder, it will exert a pressure 
against the cylinder wall over its entire surface, and thus insure 
practical gas tightness. 

Allowance for Differences in Heat Expansion—The cylin¬ 
der wall, which is in direct contact with the cooling water, never 
rises higher in temperature than the boiling point, and its aver¬ 
age temperature at full load may be taken as 200° Fahr. The 
piston head, however, which is not provided with any cooling 
means, attains a temperature of about 900° Fahr., which cor¬ 
responds to the average temperature in the combustion chamber. 
The open end of the piston will reach a temperature of about 
500° and the rings the same. The temperature of the cylinder 
walls has repeatedly been determined experimentally, but the 
temperatures of the piston and rings are based upon theoretical 
reasoning, since their experimental determination involves great 
difficulties. As regards the temperature of the piston head, 
it stands to reason that it will be slightly less than the mean 
ordinate which would be obtained if a curve were drawn with 
the cylinder temperature as ordinates and the time of a cycle as 
the abscissas, since there is little other chance for the cylinder 
head to dispose of its heat than to the cool gas of the incom¬ 
ing charge. The piston rings, which are on three sides in con¬ 
tact with metal at 700° to 8oo° Fahr. and on one side with metal 
at 200° Fahr., will naturally have a temperature about midway 
between these figures. What concerns us particularly is the dif¬ 
ferences between the temperature of the cylinder wall and the 
temperatures of the two ends of the piston and the piston rings, 
since these determine the differences in expansion that must be 
provided against. These differences are 700° for the top end of 
the piston and 300° for the bottom end and the rings. The ex¬ 
pansion of cast iron is 0.000556 inch per inch per ioo° Fahr.. 
consequently a 4 inch piston must be turned 

4x7x0.000556 = 0.0155 inch 

smaller than the bore at its head end and 

4 x 3 x 0.000556 = 0.0067 inch 

smaller than the bore at its open end. These are the allowances 
theoretically required. Slightly smaller allowances are generally 
made in American water cooled motors, probably because auto¬ 
mobile motors seldom work at full load for any length of time. 
The subject will be further discussed later on. 


110 PISTON, PISTON RINGS AND PISTON PIN. 

Forms of Piston Rings—While the piston ring bears tightly 
against the cylinder wall nearly all around there is a chance for gas 
leaks at the cut or split in the ring. Two forms of these cuts are 
in common use. These are known respectively as the diagonal 
cut and the lap joint cut, and are illustrated in Fig. 56. It is 
directly apparent that with 
the diagonal cut there is a 
leakage path for the gas if 
the ends of the ring do not 
come close together. Some 
slight space must be left be¬ 
tween the ends of the ring 
to allow for variations in 
its temperature, as com¬ 
pared with the temperature 
of the cylinder, for if the 
ends should come together 
and the ring heat up still 
more it would break. The 
aim, then, should be to 
make the cut no greater 
the difference in temperature of the ring and the cylinder 
wall is limited to 300 degrees, the heat expansion will be 
tt x d x 3 x 0.000556. * In the case of a 4 inch bore it will be 
3.14x4x3x0.000556 = 0.021 inch. 

The distance apart of the ends when the ring is compressed in 
the cylinder should therefore be slightly more than 0.021 inch in 
the circumferential direction. If the ring is cut at an angle of 
45 degrees, the actual distance apart should be slightly more than 

0.021 x sin 45 0 = 

0.021x0.707 = 0.015 inch, 

say, 1-64 inch. 

Leakage Around Piston Rings—With a lap jointed ring 
there is also a small leakage path for the gas at the joint; it is, 
however, not direct or straight, as in the case of the diagonal 
joint, but tortuous. This is illustrated in Fig. 57. We found 
* that the piston at its head end must be turned about 0.0155 
inch small in diameter, so if it were centred in the cylinder it 
would have a clearance of 0.008 inch all around. This is shown 
exaggerated in the illustration. The tongues or overlapping por¬ 
tions also do not quite fill up the slots cut for them, and the 




^ 




Z.ap Joint 

Fig. 56.— Piston Ring Joints. 


than actually required. Since 













PISTON, PISTON RINGS AND PISTON PIN. Ill 


ring at the cut is only about one-half the depth of the ring 
groove. The gas, therefore, passes from the piston clearance 
space into the lap joint clearance space, down the latter space 
to the bottom of the ring groove, thence into and up the other 
lap joint clearance and out into the piston clearance, as shown 
by arrows in the cut. 

Another thing that must be taken into account is that the 
ring, although nicely fitted, is not an absolutely tight fit in its 
groove. It has a slight amount of play in the groove, and it will 
bear against one side of the groove or the other, according to 
the direction of the resultant of the forces acting on it. The 
slight opening thus left all around between the sides of the 
piston and the ring gives plenty of chance to the gas to get 
under the ring during the compression stroke, when the ring is 
in contact with the lower side of the groove, but fortunately the 
gas can get out from underneath the ring only through the 
small joint clearance. During 
the power stroke the top rings at 
least will also bear against the 
lower side of the groove, since 
the gas pressure on them will ex¬ 
ceed the inertia and friction 
forces. 

The leakage path due to the 
joint clearance, as illustrated in 
Fig. 57, can be reduced in vari¬ 
ous ways. The lap joint clear¬ 
ance should be made as small as 
possible, so that the ends of the 
ring come almost together when 
the ring reaches its maximum 
temperature. The minimum clear¬ 
ance can be calculated in the 
same way as illustrated in con¬ 
nection with the diagonal cut ring. 

Moreover, although the piston has an average clearance of 0.008 
inch in the cylinder all around, the cuts in the rings can be so 
distributed over the circumference that at the cut in one ring 
at least the piston will always bear tightly againlst the cylinder 
wall, so there is no chance for the gas to get into the lap joint 
clearance. 

Gas tightness is particularly essential during the compression 
and expansion strokes. Now, in a motor turning right handedly 




'//////////////////yC.\ 

/tinder Watt 

1 

ipip 

!x\\\X>S\V 

Piston 
% Wall 

1 

—- 


Fig. 57.— Leakage Path 
Around Piston Ring. 





























112 PISTON, PISTON RINGS AND PISTON PIN. 


the piston is pressed 
against the right side of 
the cylinder wall by the 
connecting rod during 

the compression stroke, 

and against the left side 
of the cylinder wall dur¬ 
ing the power stroke. 

So in case four rings 
are used, the rings should be placed so that all the cuts 
will be in the plane of connecting rod travel, two on one 
side, and two on the opposite side of the piston. If only 
three are used, as is more common, two cuts might be placed 
on the left side of the piston in the case of a motor rotating right 
handedly, and one on the right side. The very common practice 
of spacing the cuts evenly over the circumference is evidently 
wrong. 

Material of Rings —In order to introduce the rings into 

their grooves they must be sprung out sufficiently to pass over 
the pistons. The material from 
which they are made must, there¬ 
fore, have considerable elasticity, 
and from this point of view a 

high carbon spring steel would be 
the most suitable. But owing 
to the fact that spring steel is 
harder than cast iron, such rings 
would tend to cut the cylinder 

bore. Cast iron is therefore used, 
in spite of its limited elasticity. 

The best grade of tough, dense 
gray iron should be chosen for 
the purpose. As a rule the cyl¬ 
indrical castings from which the 
piston rings are cut are poured 
from the same “run” in the foun¬ 
dry as the cylinders and pistons. 

It is claimed, however, that an 
iron with somewhat less silicon 
and more sulphur than the or¬ 
dinary cylinder iron is more suit¬ 
able for piston rings, being more 
springy and harder. It is appar¬ 
ent that the strains to which the material of the ring is subjected in 




JPicton 7~ravel -■— 

Fig. 58. —Position of Ring During 
Compression Stroke. 













PISTON, PISTON RINGS AND PISTON PIN. 113 


stripping it over the piston, and in compressing it in the cylinder, 
depend upon the thickness of the ring in proportion to the cyl¬ 
inder bore. The thickness must be so chosen that the safe work¬ 
ing stress of the material will not be exceeded either in stripping 
the ring over the piston, or later when the ring is compressed, 
and the ring should be so designed that the stress in the ring is 
about the same when it is being put on the piston as when it is 
in the cylinder, as under these conditions a ring exerting a cer¬ 
tain pressure per square inch against the cylinder wall will be sub¬ 
jected to the least maximum strain, and the chances of breakage 
will then be a minimum. 

Stresses in Piston Rings—In Fig. 59 is represented a sec¬ 
tion A B C D of a piston ring, the original outside radius of 
which is r, and the thickness t. Suppose that in order to strip 
this ring over the piston it has to be expanded to a radius n. 
The actual change in the radius is comparatively slight, and 
some of the dimensions in the figure are exaggerated for the 
purpose of more clearly showing the effects of expanding the 
ring. The fibres at the outer circumference of the ring will be 
compressed, and the fibres along the inner circumference of the 
ring will be extended, the section assuming the form 
A', B', C', D’ . It is shown in text books on the mechanics of 
materials that the relative extension of material under tension 
and its compression under pressure are equal to the stress S, to 
which it is subjected, divided by its coefficient of elasticity E. 
Therefore, referring to Fig. 59, let N be the tension on the inner 
fibres of the ring, and —N the compression on the outer fibres. 
The tension and compression are equal, because the ring is sub¬ 
jected to plain bending stresses. We may then write 


— .S’ 

A' B' — A B 

A' B' _ 

E “ 

A B 

~ AB 

5 

D' C' —DC 

d' a _ 

E ~ 

D C 

~ DC 


Transposing, 

A' B' — S 
A B “ E 


-fi. 


D'C . 
D C — £ 


( 34 ) 

( 35 ) 


But A B : D C — r : r — t , 


and A' B' : D' C' = r 1 :r 1 - • 
Dividing the lower proportion by the upper 

A'B' D'C' r x r x — 1 . 












114 PISTON, PISTON RINGS AND PISTON PIN. 


A' B' D' C f 

Substituting the values of — 7 - 7r and found in (34) and (35), 


A B 


- S 1 *5" I A | 

+ 1 • T . + 1 = - • 

E E r 


D C 

>'1 — t 


n 

r 


E 


>'1 ■ 

— t 

( S - 

r - 

- t ~~ 

u 


3 



A \r 

+ r- 


— t 

J_\ 

r 

— t 

r 

n 

1 _ 

- t~ 

r 

+ r _ 

-t 


— — (r —/) _ 

*\ (r — l) H- r (r x — /) 


t (r* 


r) 


2 r r t — l (r + r x ) 


Consequently 

t (r, — r) 


S — E 


2 r r x — t (r + r,) 


( 36 ) 



This equation gives the stress in the outermost fibres of the 
ring when expanded to an outer radius n to be stripped over the 
piston. It is now to be determined what this radius r x must be. 

If the ring is to be subjected to the same stress when stripped 
over the piston as when compressed in the cylinder, its outer 
radius r when in 
the free state must 
be a mean between 
the radius n when 
it is expanded, and 
the radius r 2 of 
the cylinder bore. 

Now, suppose that 
the ring is concen¬ 
tric, of thickness t, 
and that it can be 
,so manipulated that 
it will form a true 
circle when ex¬ 
panded so that it 
will just pass over 
the piston. Then the outside radius n will be r 2 + f. In reality the 
ring when expanded by applying opposing tangential forces to its 
end, as shown in Fig. 60. will not form a true circle, but approach 


Fig. 60. 






















PISTON, PISTON RINGS AND PISTON PIN. 115 


an ellipse in form, owing to the fact that when thus expanded the 
bending moment on the ring is a maximum opposite the cut, 
and that is where most of the bending takes place. This ten¬ 
dency to localization of the bending can be counteracted by ap¬ 
plying pressure to the ring in the direction toward the centre, 
about 75 degrees on each side of the cut. By applying this ex¬ 
pedient when inserting the ring into its groove it need not be 
expanded to an appreciably greater radius than r 2 + t to get it 
over the piston. We then have 

r \ ~ + i 

i t 

r = r 2 + - 
2 

Inserting these values in equation (36) 


S = 


2 r. 




t l 7 

4 r 2 2 - - + - 
4 


2 ^ 
2 


But r 2 is equal to one-half the cylinder bore, or to — . Hence 

2 


t 2 



In this discussion we have assumed that the ring is concentric, 
i. e., of uniform thickness all around, and that when it is stripped 
over the piston its curvature is changed uniformly over the 
whole circumference. In practice, however, these conditions do not 
entirely hold true. Both in expanding the ring and in compress¬ 
ing it the stress will be greatest at the point opposite the cut, 
and it is, therefore, this part of the ring that demands special 
investigation, because if that is so proportioned that it is not 
overtaxed, the rest of the ring will be safe. 

In view of the fact that t in actual piston rings is only about 


— > the expression — _i_— in the denominator of the right hand 
30 4 4 

member in equation (37) can be neglected in comparison with 
tf without appreciable error (less than one per cent.), and the 
equation then becomes 


(38) 








116 PISTON, PISTON RINGS AND PISTON PIN. 


which may be transformed into 



The coefficient of elasticity E of cast iron is about 15,000,000. 
Since the stress is practically unchangeable, it may be compara¬ 
tively high, say 15,000 pounds per square inch. These values in¬ 
serted in equation (39) give 


t 



15,000 

15,000,000 


b 


b 

32 


(40) 


With a concentric ring this is nearly the maximum thickness 
which can be safely gotten on to the piston, though with an ec¬ 
centric a somewhat greater thickness is permissible at the point 
opposite the cut. 

The general practice in this country is to have the ring one- 
half as thick at the cut as opposite the cut. In that case, calling 
t the thickness opposite the cut, we have 


r x = r 2 -f | / 
r = r 2 + f t 


which inserted in equation (36) give 


S = E 


3 /2 
? I 


2 *-,« - A t' + r *‘ 


r, t 


Neglecting the expression —-* 

4 

r, _ Q / 2 

S—E D 


16 r 2 2 


Substituting - for r 2 we get 
2 


? 2 
4 d* 


s=eA-L , 


this reduces to 


and assuming the same stress and coefficient of elasticity as 
above, we get for the permissible thickness opposite the cut 


*=V 4 -|=— 

r a A 27. 


27.5 


(41) 


This gives about 5-32 inch for a 4 inch piston, and i 3 s inch for 
a 5 inch piston. 

Pressure Against Cylinder Wall—This pressure is radially 
outward and should preferably be uniform over the entire 
surface of the ring. It is balanced by an equal and opposite 
reaction of the cylinder wall against the ring. We will denote 
this reaction per square inch by p, and the width of the ring 
(the dimension in the direction of piston travel) by w. 











PISTON, PISTON RINGS AND PISTON PIN. 117 

The reaction of the cylinder wall on the ring produces a 
bending effect in the latter which subjects the material -of the 
ring to a certain stress. The pressure of the ring against the 
cylinder wall and the stress in the material of the ring are 
mutually related, and we will now investigate this relation. 
We will take a cross section of the ring at C, Fig. 61, and find 
the bending moment for that section. 



Consider a small strip da of the ring surface between A and C. 
The area of this strip is evidently w da, and the total pressure 
on it, p w da. This pressure can be divided into two com¬ 
ponents, one, p w da sin <P, parallel to the cord A C, and the 
other, p iv da cos <f>, perpendicular to the cord A C. It is di¬ 
rectly apparent that, considering the pressure acting on the 
mtire arc A C, the components parallel to cord A C vanish. 




















118 PISTON, PISTON RINGS AND PISTON PIN. 

and it is only the components perpendicular to AC that pro¬ 
duce a bending moment at C. 

Considering the expression for this perpendicular component, 
we find that 

da cos 0 = dc t 

because 

Angle G E F (0) = angle H E M, 

H E being perpendicular to E F and E M perpendicular to 
E G, and 

Angle H E M = angle E M C, 

H E being parallel to M C. The line EM is a tangent to the 
arc A C at E, and comprises, therefore, the short section da of 
this arc. 

The component of the radial pressure on da perpendicular to 
A C is therefore p zv dc. That is, that component of the radial 
pressure on any small strip da across the circumference of the 
ring which produces a bending moment at C is equal to the 
product of the area of the projection of that strip on the cord. 
A C, into the radial pressure. As this applies to every small 
strip of the arc A C, it follows that the total pressure producing 
a bending moment at C is 

2 p zv dc = p zv A C. 

But 

a r a r\ ' ® $ 

A C= 2 A O sin — = 2 r 2 sin — • 

2 2 

The above force acts on the point C through lever arms vary- 

ing in length from zero to 2 r 2 sin — , the mean length being 

2 

; • 6 
l— r 2 sin — » 

2 

consequently the bending moment at C is 

71 f _ / 2*2^ 0 . n 0 

M =■ 2 p zv r 2 sin'* — — -sin 2 _• 

2 2 2 

The resistance to bending of a rectangular section of base zv 
and height t is 

Mr= — = ■ 

c 6 

The bending moment is equal to the resisting moment, and 
we may therefore write 

zv b 2 . 2 8 S zv A 
- - sin 2 — — -? 

2 2 6 

from which it follows that the radial pressure of the ring is 
S F 









PISTON, PISTON RINGS AND PISTON PIN. 119 


We started out with .the supposition that p was constant, 
and by analyzing equation (41) we find that in case the 
thickness of the ring is constant, p varies inversely as the 
square of the sine of half the angular distance from the cut 
in the ring. The only way in which p can be made constant is 
by varying the thickness t of the ring with the angular distance 

d 

from the cut, t appearing in the numerator and sin — in the 

2 

denominator of the expression for p. Equation (41) may be 
transformed to read. 


t — b 



For the point opposite the cut, for which 


sin 


6 , 

— = 1, we have 
2 



— =90 degrees and 

2 


But the thickness at this point has already been fixed by equa 
tion (39), which may be changed to read 



Hence we may write 



E ~ S 



(42) 


This equation shows that the radial pressure of the ring 
varies only as the square of the stress in the material; hence, 
since the permissible stress is limited, the radial pressure is 
closely limited as to its higher values. If we assume the former 
values for the working stress and coefficient of elasticity, viz., 
S= 15,000 pounds per square inch and £ = 15,000,000, then 


p — JS' 000 X 15,00 0 _ ^ -pounds per square inch. 

3 X 15,000,000 

Eccentric vs. Concentric Rings—At 90 degrees from the 
cut, for which 

— = 45 degrees, 








120 


PISTON, PISTON RINGS AND PISTON PIN. 


and 


sin 1 ' — — 0.5, 

2 

the thickness should be 

*90*= bA/2J! = t*S \ = 0.707 t. 

At the ends of the ring, that is, at the cut, for which 

angle— = O, 

2 


. j e 

sin 2 — = O, 

2 

and the thickness of the ring should therefore be zero. This 
condition must be fulfilled if a ring which is originally turned 
on its outer circumference to a circle of greater diameter than 
the bore is to fit into the bore with a uniform pressure over 
its entire surface. In practice, however, it would be impossible 
to reduce the thickness of the ring to nothing at the cut. 
Not only would such a ring be exceedingly delicate in handling, 
but it would certainly not be gas tight near the cut. The 
above mathematical discussion, however, clearly shows the reason 
for making piston rings eccentric, that is, of less thickness 
near the cut than opposite the cut. This is done with the object 
of making the pressure on the cylinder wall more nearly equal 
all around the ring. Eccentric rings are almost universally 
used in this country, though in Europe concentric (uniform 
thickness rings) are extensively used. A noted German writer, 
Hugo Giildner, in his work on the internal combustion motor, 
recommends the use of concentric rings, because, as he says, the 
object of eccentric rings, viz., uniform pressure all around, 
cannot be attained anyhow. While an absolutely uniform pres¬ 
sure cannot be attained in this way, uniformity can be approached, 
and that is worth something. Of course, this advantage is 
partly offset by the fact that the eccentric ring leaves con¬ 
siderable space in the groove near the cut and the gas will pass 
around it easier if it does not make a tight joint with the 
side of the groove. 

Width of Ring —The consensus of opinion among engineers 
is that it is best to make the rings comparatively narrow. For 
instance, for a 5 inch piston, four rings each 9-32 inch wide, 
would be considered superior to three rings each inch wide. 
A good rule is to make the width of the ring 


iv = 


b 

20 


cocccoo^eo*** 


.(43) 



PISTON, PISTON RINGS AND PISTON PIN. 121 


If the rings are made narrower the length of the piston can 
be reduced and the weight of the reciprocating parts thus minim¬ 
ized, which increases the speed of maximum output, and con¬ 
sequently the output itself. A shorter piston also means a 
shorter cylinder, and consequently a lighter motor. Besides, 
the amount of power lost on account of the friction of the 
rings increases almost directly with the width of the rings. If 
the rings are carefully made so they bear firmly against the 
cylinder wall all around, nothing will be gained by making the 
width of the rings more than one and one-half times their 
maximum thickness. 

Manufacture of Rings—Piston rings are made from hollow 
cylindrical castings or blanks, Fig. 62, cast with three or four 



Fig. 62.— Piston Ring Blank. 

lugs at one end, by means of which they can be bolted to the 
face plate of a lathe or to a special plate, which in turn is bolted 
to the face plate. The castings are generally made of sufficient 
length to furnish from 12 to 15 rings each. As a rule, the 
blanks are turned up on the outside and bored out on the in¬ 
side, but in some cases they are not finished on the inside, be¬ 
cause it is considered that the scale makes the ring more springy. 
At the end where the lugs are the casting may be provided with 
a flange to stiffen it while it is being machined. 

A fixture for turning the ring blanks eccentrically is shown 
in Fig. 63. A circular disc A is secured to the face plate B 
of the lathe by means of three bolts. It swivels around the top 
bolt C, this being permitted by oblong slots through which the 
other two bolts D and E pass. The swivel motion is limited by a 
stop pin F. The ring blank is bolted to the plate A by means, 
of four studs In machining up the casting, the disc A is first 
swung over to one side as far as stop pin F permits, and fixed in 
place, and the outside is turned up. Then the disc is swung 
over to the opposite side and tightened, and the casting is bored 













122 PISTON, PISTON RINGS AND PISTON PIN. 




Fig. 63.— Eccentric Turning Fixture. 


out. The degree of eccentricity depends, of course, upon the 
range of travel of the pin F in its slot. In the illustration, the 
distance from C to the centre of the casting is one-half the dis¬ 
tance from C to F, and the eccentricity will therefore be equal 
to one-half the motion of F in its slot. 

After the blank is thus finished up, the rings are cut at the 
same setting by means of a multiple cutter tool, as shown in 
Fig. 64. A casting A, secured to the lathe carriage, is drilled 
with a number of holes equal to the number of rings to be cut 
from one casting. These holes are tapped for adjusting screws 
for the tools, and vertical slots are then milled in the casting 
for the cutters, central with the adjusting screw holes. The cut¬ 
ters are clamped down to the tool rest in pairs by means of small 
clamping plates held by two screws each. The tools are set so 
that any one of them has a slight lead over the one next to it in 
the direction toward the lathe head, so that the ring cut from 



Fig. 64.— Cutting Off Tool. 


























































































PISTON, PISTON RINGS AND PISTON PIN. 123 


the end of the casting farthest from the head will be detached 
first. Since the casting remains in the position in which it 
was bored, it will revolve concentrically with the inner surface, 
which will facilitate the cutting off operation. It is common to 
slightly round the nose of the tools on the side toward the lathe 
head. This leaves a fin on the casting when a ring is detached, 
which is removed before the next ring is cut off. 

Grinding Rings—It is now common to grind piston rings on 
both sides and on the circumference, and in turning the blank 
and cutting off the rings the proper allowance must be made for 
this grinding. The allowance in width for grinding on the sides 
should be from twenty to thirty-one thousandths of an inch. The 
next operation consists in grinding the sides of the ring. This is 
done on a special piston ring grinding machine on which the 



Fig. 65. —Heald Piston Ring Grinder. 






124 PISTON, PISTON RINGS AND PISTON PIN. 


rings are held by means of a magnetic chuck. Such a piston 
ring grinder is illustrated in Fig. 65. 

Next the rings are cut open. This operation differs, of course, 
according to whether the ring is to have a diagonal joint or a 
lap joint. In either case the cutting is done on a hand miller 
with a cutter of exactly the right width for the slot to be cut. 
After this operation it is best to submit the rings to a test before 
doing any further work upon them, so as to minimize the waste 
due to rejections. The easiest way to make this test con¬ 
sists in forcing the ring over a tapered arbor to a certain 
mark and rejecting all rings which show more than a certain 
permanent set, determined by measuring the width of the slot 
before and after the test. 

Next the rings are ground on the outside, and in order that 
they may perfectly fit the cylinder after being ground and exert 



an even pressure all around their circumference, as nearly as that 
is possible, they are first compressed in a fixture, then clamped 
on an arbor while in the compressed state, and finally ground • 
down to the exact diameter of the cylinder bore. A clamp suit¬ 
able for compressing the rings is illustrated in Fig. 66. It con¬ 
sists of two semi-cylindrical shells of cast iron with double 
hinges on one side and a hinged clamp screw and correspond¬ 
ing lugs on the other. This clamp is bored out to a diameter 
equal to the bore, plus the grinding allowance. The rings are 
assembled into this clamp in such a way that the cuts are stag¬ 
gered, which makes them hold together better after the clamp 
is removed. This staggering is facilitated by three slots cut in 
the inner surface of the clamp at equal angular distances. When 
the rings are all assembled in the clamp, the latter is drawn up 
tight and is passed over the arbor, Fig. 67, on which the rings 
are to be ground. This arbor is provided with two collars. 
The one on the left is of such a diameter that the clamping 














PISTON, PISTON RINGS AND PISTON PIN. 


125 



Fig. 67. —Ring Grinding Arbor. 


fixture will just pass over it, by which means the rings are 
centred on the arbor. The collar on the right hand side is of 
the exact diameter of the cylinder bore to which size the rings 
are to be finished. The collar on the left may be integraDwith 
the arbor, but if it is loose it should be held from turning by a 
pin or key, as should the other collar. This collar is drawn up 
tightly by means of a hexagonal nut, the clamp is then removed, 
the arbor is placed in the lathe or grinding machine and the 
rings are ground to size. While one set is thus being ground the 
clamp can be filled up again with rings. The collars should 
preferably be case hardened. 

Hand Fitting of Rings —The above description applies to 
the method of manufacturing piston rings in large plants. When 
rings are made on a small scale, as for experimental motors, it 
is, of course, impossible to install the somewhat elaborate equip¬ 
ment described above, and the rings are then fitted by hand. They 
are cut off from the blank with a tool which gives a very smooth 
surface, the feed being comparatively slow, so that the sides will 
fit a carefully turned groove without any further finishing.- After 
the ring has been split by means of a saw or file (the latter in 
the case of the lap joint ring), the workman will file the ring 
off on the outside at those points where he knows from experience 
that metal has to be removed in order to make the ring fit the 
bore. This applies particularly to the ends of the ring near the 
cut. Next, the surface of the bore is smeared with a mixture of 
red lead and oil, the ring is introduced into the bore, then with¬ 
drawn, and any spots which are reddened are filed down more, 
because the radius at these points must evidently be reduced in 
order to make those parts of the ring which are not reddened 
bear against the cylinder w r all. In this way, after repeated trials, 
a fair fit of the ring in the cylinder is finally obtained. The ring 
can be further fitted after the motor has been run for some time, 
when it will be noticed upon removing the piston from the 









































126 PISTON, PISTON RINGS AND PISTON PIN. 


cylinder that the rings are polished at some points and blackened 
with carbon at others, where the gases have blown through. 
By filing down the polished portions of the ring, the degree of 
gas tightness can be further increased. 

Peening the Rings—In foreign shops piston rings are often 
fitted by peening. This operation consists in striking the surface 
of the ring sharp blows with a special peening hammer while it 
is being held on an anvil. Successive blows are struck at short 
distances until the entire portion of the ring which seems to have 
too big or too small a curvature has been covered. The blows 
from the peening hammer spread the metal, consequently by 
peening the ring on the outside the radius of curvature of that 
portion of the ring which is being peened is reduced, while by 
peening the ring on the inside the radius of curvature is in¬ 
creased. The method is illustrated in Fig. 68. 



Fig. 68.—Peening Piston Rings. 


Dimensions of Ring Blanks—We will assume that the ring 

is to be eccentric, that its maximum thickness is to be —— 

27 ‘ 5 

and its minimum thickness - . Then the outside diameter 

55 

of the finished ring in the free state is evidently 












PISTON, PISTON RINGS AND PISTON PIN. 127 


b -f- 2 J-ll -55^ _ j. # Q2 y 

2 

For grinding, an allowance of 0.008 b should be made. Hence 
the blank should be turned to an outside diameter 


do — i. 027 b o. 008 & = 1.035 &.(44) 

and to an inside diameter 

di—b — 0.027 b = 0.973 h . ( 45 ) 


In splitting the ring, enough stock must be taken out so that 

when the ends are brought to within inch of each other, 

250 

the new diameter is equal to the bore plus the finish allowance, 
that is, 

b + 0.008 b = 1.008 b. 

The outer circumference must therefore be reduced by 
1.035 7r b — 1.008 7 T b = 0.027 7T b = 0.085 b inch. 

The slots cut are equal in width to this reduction in the circum¬ 
ference plus the clearance allowance, which is — or 0.004 b, 

250 

making a total of 0.089 b, say 0.09 b. 

The above rules for dimensioning the rings are based on a 
stress of 15,000 pounds per square inch in the metal when the 
ring is compressed in the bore and slightly more when it is 
stripped over the piston (because of the impossibility of expand¬ 
ing it so its inner circumference will just fit the piston). If the iron 
from which the rings are made will not stand this stress without 
taking an undue permanent set, then the formulae can be changed 
on the basis of a lower stress. But it is desirable to have the 
rings as stiff as the material will stand, as then the cylinder 
will be more nearly gas tight and there will be less trouble 
from carbon adhering to the finished surface of the bore. 

The reason why quite a few manufacturers favor the diagonal 
joint, when the lap joint is generally considered to be better from 
the standpoint of gas tightness, is undoubtedly that the tongues 
of a lap jointed ring for an automobile motor aie very small 
and correspondingly delicate. For a 4 inch bore, for instance, 
the tongues are only about 3-32 inch wide. In order to make 
them less breakable they should be provided with a fillet, as 
shown in Fig. 69. It is also advisable to have the tongues some¬ 
what shorter than the length that must be cut out of the ring in 
order to reduce its diameter sufficiently. For instance, it was 
stated above that practically Y inch must be cut out of a 4 inch 
ring to allow it to be compressed from the original diameter 
to the diameter of the bore. But a 14 inch tongue would be 






128 PISTON, PISTON RINGS AND PISTON PIN. 



Fig. 69.— Slotting Lap Joint Ring. 

easily long enough, and the two slots in the ring must then be 
milled so that they overlap y$ inch. After milling, the inner faces 
of the tongues, which are to go together, must be dressed up 
with a fine file and before the ring is finally passed on to the 
stock room one of the things that should be ascertained by the 
inspector is that when compressed to an outside diameter equal 
to the bore there is a clearance at the end of the tongues equal 
to 0.004 b. 

Diagonal cut rings are most extensively used at present for 
automobile motors. The angle made by the sides of the cut with 
the sides of the ring varies between 30 and 45 degrees. The 
amount to be cut out of the ring in the circumferential direction 
is, of course, the same as in the case of the lap jointed ring, 
viz., 0.09 b. But the milling cutter used in cutting the slot 
should be only 0.09 b sin a in width; that is, for a 45 degree slot 
cutter should be 0.064 b wide and for a 30 degree slot 0.045 b. 

Some manufacturers, in order to insure that the top ring will 
bear firmly against the cylinder wall, cut a couple of small slots 
in its top side, so the pressure in the combustion chamber will 
get into the groove and press the ring out. This has the advan¬ 
tage that the ring pressure against the cylinder wall will be 



Fig. 70.—Marmon Compound Piston Ring. 


























PISTON, PISTON RINGS AND PISTON PIN. 129 

greatest when gas tightness is most essential, but it adds to the 
friction and wear. 

Special Rings—A number of manufacturers have adopted 
special designs of piston rings, with the object of preventing 
the leakage due to the cuts in the ordinary ring. A ring of this 
type, the Marmon, is illustrated in Fig. 70. It consists of one 
wide concentric ring which goes into the bottom of the groove, 
and two narrow concentric rings which are placed in the top of 
the groove. All three of the rings are slotted straight across, 
and the three cuts are spaced 120 degrees apart, the narrow 
rings being held in position on the wide ring by pins riveted 
into the latter. The advantage of this compound ring is that the 
gas cannot blow through the joints, owing to the fact that the 
joints in adjacent rings are located at an angular distance of 
120 degrees and the narrow rings fit nicely together on their ad¬ 
jacent sides, where they are ground, and are pressed against on 
their under side by the wide ring. Two such compound rings 
are used on each piston. 




Fig. 71.—Special Forms of Piston Rings. 


Considerable ingenuity has been exercised in recent years in 
the elimination of the gap in the ring, which is one of the 
causes of leakage. Abroad a ring comprising two complete 
turns, known as the Lehmann double spire ring (Fig. 71 A), 
seems to have considerable vogue. It is a concentric ring and 
apparently is cut with a very fine circular saw r after the blank 
has been turned up and cut off. In this country the leakproof 
piston ring, shown in Fig. 71 at B, has been on the market for 
some years. It consists of two separate rings, one having an 
outward and the other an inward flange, the flange on one ring 
obturating the slot in the other. Both rings are turned concen¬ 
trically, and uniform pressure is aimed at by placing the slots 
on opposite sides, so that the non-uniformity of the outward 
pressure of one ring is compensated by that of the other. The 
separate rings are prevented from sliding upon one another by 
a pin, but the complete ring is free to turn in the groove. 










130 PISTON, PISTON RINGS AND PISTON PIN. 


Pinning Rings—If the rings are to remain in a certain 
position in the grooves, they must be prevented from rotating 
by means of pins. In large gas engines the rings are always 
pinned, and the same practice was formerly followed in the 
automobile industry, but, owing to the fact that pinning of the 
very small rings required for automobile motors involves con¬ 
siderable difficulty, it has been largely given up. The argument 
in favor of pinning is that if the rings are free they may rotate 
in their grooves, and it may happen that all of the cuts come in 
line with each other, in which case the gas tightness would be 
seriously impaired. Automobile manufacturers, however, have 
reached the conclusion that the chances of all of the cuts coming 
in line are exceedingly small, and even in that case there will 
not be such serious leakage if the joint clearance is small. In 
heavy gas engine practice it is customary to rivet a pin into 
the piston wall at the middle of the groove and drill a hole 
through the ring at the middle of its width, into which the pin 
may extend. But with rings as narrow as commonly used for 
automobile motors, even a one-sixteenth inch hole would seri¬ 
ously weaken the ring, by concentrating the stress. Conse¬ 
quently, about the only practical method of pinning a ring con¬ 
sists in drilling two small holes through the ring at that point 
where the slot is to come, at a distance apart equal to the 
width of the slot, so that when the ring is slotted and the ends 
are brought together there will be a space at the joint for the 
pin. There is undoubtedly less need for pinning the rings when 
both cylinder and rings are ground truly cylindrical than when 
they are fitted to each other by a prolonged “running in” process. 

Handling Piston Rings—It has already been pointed out 
that piston rings, especially for such small cylinders as are used 
in automobile work, are extremely delicate and must be handled 
very carefully when placed on the piston, and still more so when 
taken off. In expanding the ring previous to stripping it over 
the piston, pressure must be applied to the free ends evenly, 
and various simple tools for this purpose have been devised. 

The Piston Pin—The form and dimensions of the piston 
are determined by both the piston rings and the piston pin, and, 
having concluded our study of the piston rings, it will be advis¬ 
able to take up the piston pin before passing on to the piston 
itself. The piston pin is a cylindrical metal part which is 
generally firmly secured at its ends in bosses cast integral with 
the piston, and whose central portion has a bearing in the small 
end of the connecting rod. Sometimes the arrangement is re- 


PISTON, PISTON RINGS AND PISTON PIN. 131 


versed and the piston pin is clamped in the small end of the 
connecting rod and has bearings in the piston bosses. The piston 
pin dimensions are calculated by considering the pin first as a 
bearing journal carrying a certain load per square inch of pro¬ 
jected area, and, secondly, as a beam supported at both ends 
and uniformly loaded between supports. A tubular pin is pref¬ 
erable to a solid one, because for the same maximum fibre 
strain and bearing pressure it will be lighter. Hence, the piston 
pin is generally made hollow. 

Maximum Explosion Pressures—In determining the proper 
diameter of the piston pin (as well as other parts subjected to 
the force of explosion) account must be taken of the fact that 
widely different compression pressures are in use, and that in 
consequence the explosion pressures in different automobiles dif¬ 
fer greatly. Of modern stock car motors some use a compres¬ 
sion pressure of 60 pounds per square inch gauge and others as 
high as 90 pounds. The maximum explosion pressures are 
roughly proportional to these figures, and it is obvious that if 
the explosion pressure in one motor is 50 per cent, greater than 
in another, the former should have relatively stronger parts. 

The maximum explosion pressure is a somewhat elusive figure, 
for it depends not only on the compression ratio but also on the 
form of the compression chamber, the location of the spark 
plug terminals, the intensity of the spark, the temperature of the 
cylinder walls and possibly other factors. The maximum pressure 
undoubtedly occurs when, owing to defective ignition, several 
charges have been missed and the motor speed has been reduced, 
in consequence of which the compression space contains an un¬ 
usually large charge of pure mixture. It is obvious that no 
very accurate figure can be given for the maximum explosion 
pressure, but we will not be far wrong in assuming that this 


TABLE IV. 


Maximum Explosion 

Compression Pressure (/>), 

Ratio (r). (Lbs. P. Sq. In., Gauge) 

3 230 

3-2 250 

3-4 274 

3-6 298 

3- 8 321 

4 344 

4*2 368 

4- 4 392 

4-6 414 

4-8 437 

5 . 460 














132 PISTON, PISTON RINGS AND PISTON PIN. 

pressure is four times the absolute compression pressure based 
on the assumption that the cylinder is filled with charge to 
atmospheric pressure at the beginning of the compression stroke. 
This gives maximum pressures for different compression ratios, 
as compiled in Table IV. 


Calculation of Piston Pin—The length of the bearing por¬ 
tion of the piston pin should be as great as possible, but it is 
limited by the consideration that if the pin is fixed in the piston 
bosses the latter must be sufficiently long to provide room for 
a means for locking the pin, such as a set screw. In automo¬ 
bile motors the maximum length available for the bearing is 
equal to about one-half the bore. Therefore 

i = —.(46) 

2 

With a maximum explosion pressure p per square inch, as de¬ 
termined from the above table, the total pressure on the piston is 

p=z£±. 

4 

When the motor is being started or is running at very low 
speed the pressure on the piston pin is the same as that on the 
piston. At high speed the former is considerably less than the 
latter, owing to the inertia of the piston, which opposes the 
force of the explosion. But the size of the piston pin should be 
governed by the maximum pressure sustained by it, and the 
inertia may therefore be neglected. 

If we denote the diameter of the pin by d, then the projected 
bearing surface is 


s = d l = 


db 




2 

and the maximum unit bearing pressure is 


7 r b 2 p 



db 2 d 


2 

The average pressure on piston pin bearings is about 2,500 
pounds per square inch. This may seem very high, but is per¬ 
missible because the rubbing speed is so low. It, however, neces¬ 
sitates the use of a hard bearing metal lining, such as phosphor 

bronze, in the connecting rod. 

Hence 





PISTON, PISTON RINGS AND PISTON PIN. 133 


<* = —(appr.). (47) 

The inside diameter of the pin may be determined as follows: 
The pin is one of the reciprocating parts and should be as light as 
possible—consistent with the necessary strength. It is, there¬ 
fore, generally made of a material of high tensile strength, such 
as high carbon steel, nickel steel or nickel-chrome steel. If 
3 y 2 per cent, nickel or an equally tenacious steel is used a work¬ 
ing stress of 20,000 pounds per square inch may safely be allowed. 

The piston pin may be regarded as a uniformly loaded beam 



Fig. 72. 


supported at both ends (Fig. 72). It is made a pretty tight fit 
in the bosses, and the latter are rendered rigid by a wide rib 
connecting them to the cylinder head. Calling P the total pressure 
on the pin and / its length between supports, the maximum bend¬ 
ing moment is 

PI Pb 
8 16 64 

The resisting moment of a hollow section of outside diameter d 
and inside diameter d i is 



Equating the bending and resisting moments, 

7r b 3 f St Id K — cf 4 i\ 

64 _ 3 2 \ d ) 

= d' —rfi* 

2 S 

2 6’ 








































134 PISTON, PISTON RINGS AND PISTON PIN. 


and 



We will now calculate the sizes for a piston pin for a 4x5 inch 
motor with a compression ratio of 4.5. The maximum explosion 
pressure with this compression ratio is 400 pounds per square 
inch. The length of bearing is 

= 2 inches. 

2 

The outside diameter of the pin is 

= 4 X 400 = t inch ' 

1,600 1,600 

The inside diameter is 

\/ i 4 — 1 ^ ^- 4 _ QO = 0.77 inch, 

r 2 X 20,000 

say 24 inch. 

If ordinary carbon steel is used it is better to limit the stress 
to 15,000 pounds per square inch, and the inside diameter would 
then work out to about §4 inch in the above example. It must 
be remembered that the piston pin works under rapidly alternat¬ 
ing stress; that is, under the most severe conditions possible. 
The ends of the bore through the pin are frequently tapered 
out, since there is comparatively little strain on the pin there. 

Piston Bosses—The diameter of the piston bosses must be 
about 1.5 d, where d is the outside diameter of the piston pin, 
and as a rule the bosses are slightly tapered, increasing in diam¬ 
eter toward the piston wall. 

Locking Piston Pins—The piston pin must be prevented from 
moving lengthwise in the piston bosses or “drifting.” for if one 
of its ends should come in contact with the cylinder it would 
soon start to cut the cylinder bore, especially if the pin is case 
hardened, as it frequently is. In order to minimize the cutting 
if the pin actually should come loose, it is well to round off 
the outer edges of the pin when it is being machined. 

One method of holding the pin in place lengthwise consists 
in placing a piston ring around the piston central with the 
piston pin, as shown at A, Fig. 73. This ring is usually made of 
somewhat greater width than the regular piston rings, but of 
the same maximum thickness. A second method, which is em¬ 
ployed on quite a number of foreign motors, is illustrated at B . 
A hole is drilled through the wall of the hollow piston pin to 









PISTON, PISTON RINGS AND PISTON PIN. 135 


receive the tapered end of a pin screw screwed through the wall 
of the piston boss. The end of the piston pin is split through 
the centre of the hole for the locking screw with a saw, so that 
the pin will be expanded by the tapering portion of the pin 
screw and securely held in the piston boss. The pin screw is 
prevented from unscrewing by means of a split pin, the pro¬ 
jecting ends of which are longer than the distance of the screw 



Fig. 73.—Piston Pin Locking Means. 

from the piston wall. A somewhat similar method is shown 
at C. In this case the pin of the locking screw is not tapered, 
and instead of the split pin through the head of the screw, a 
nail or pin is passed through a hole in the pin of the pin screw 
inside the hollow piston pin. 

At D, Fig. 73, is shown a patented locking device used on the 
motors of the Autocar Company. That portion of the piston 
pin where the locking device comes is left solid, and a radial 
hole is drilled into it into which is inserted first a coiled spring 
and then a double diameter plunger. When the piston pin is put 
in place this plunger is “submerged” in the hole, and when the 
plunger comes opposite a hole drilled in the wall of the piston 
boss, the outer, smaller diameter portion of the plunger will enter 



















































































































136 PISTON, PISTON RINGS AND PISTON PIN. 

this hole, and thus lock the piston pin in place. At E is shown 
the locking method employed by the Locomobile Company of 
America. This is very similar to the method illustrated at C, 
except that in place of the pin used at C a wire is passed 
through the holes through the two pin screws and has its ends 
turned back. At F is shown the locking means employed on the 
Franklin motor. The piston pin is held from rotating by a pin 
passing transversely through one of its ends and into a slot cut 
in the wall of the piston boss from the outside of the piston; it 
is held against longitudinal motion by a split pin passing through 
the piston boss and piston pin wall at the other side of the 
piston. 

Piston Pin Fastened in Connecting Rod. —Quite a few 
manufacturers now secure the piston pin in the connecting rod 
small end, so that the pin has bearings in the piston bosses. One 
advantage of the arrangement is that larger bearing surfaces 
can be secured in this way than with the pin bearing in the con¬ 
necting rod. Among the American makers which follow or have 
followed this practice may be mentioned Columbia, Selden, Amer¬ 
ican, Speedwell and Midland. The length of the bearing surface 
can be made as much as 0.7 b with this construction. The 
outside diameter may be found by means of equation (47). In 
order to find the required inside diameter we proceed as fol¬ 
lows : The ends of the pin in this case form cantilever beams 

uniformly loaded, the total load on each being -. If we con¬ 


2 


sider that the projecting end of the pin is 0.35 b then the maxi¬ 
mum bending moment is evidently 


— X 0.35 b 
2 


— 0.0875 Pb = 


— 0.0875 


0.0875 7 rtf ' f t 


2 

The resisting moment is again 


4 


S 7 r 
32 




Therefore 











PISTON, PISTON RINGS AND PISTON PIN. 137 


df-^d'- 0 -1*1* 

S 

When the piston pin has bearings in the piston bosses it is 
generally secured in the connecting rod by means of clamping 
screw which passes through a split boss on the side of the 
hub at the upper end of the rod. The hole for the screw 
is drilled in such a position that the screw will pass slightly 
beneath the surface of the piston pin, the latter being grooved 
for the purpose. The screw may be locked in position by means 
of a spring washer, but in order to insure absolute security it 
is well to add a split pin, passed through the head of the screw. 



of such a size that the connecting rod will prevent the screw 
from turning. The most common arrangement is illustrated in 
Fig. 74 and a variation is shown in Fig. 75. 

Where the piston pin has its bearings in the piston pin bosses 
it is not necessary that the latter should be provided with 
bearing bushings, as cast iron is a good bearing material under 
low-speed operating conditions, and it is customary to have 
the pin bear on the bosses direct. This makes it possible to 
reduce the weight of the piston pin and of the piston, and, of 
course, is advantageous also on account of the saving in labor 
it effects. 

Side Pressure on Cylinder Wall—It was shown in a pre¬ 
vious chapter that when the momentary effective pressure on the 
piston per square inch of piston head area is denoted by /p and 
the angle which the connecting rod makes with the cylinder 
axis by <f>, then the reaction of the cylinder wall against the 
piston per square inch of piston head area is 








































138 PISTON, PISTON RINGS AND PISTON PIN. 


P % — P p tan 0 . (Equation 25A). 

The pressure of the piston against the cylinder wall is equal and 
opposite, and for practical purposes may be represented by the 
same equation. It is then necessary to find the values of the re¬ 
sultant pressure and of tangent 0 for different points of the 
stroke and multiply corresponding values together. The effective 
piston pressure is equal to the resultant of the momentary gas 
pressure in the cylinder and the inertia force. It will be con¬ 
venient to take points corresponding to the completion of 0.1, 
0.2, 0.4, 0.6 and 0.8 of the stroke, beginning at the top end, and 
determine the values of the resultant pressure and of tan 0 for 
these points. We will take the same indicator diagram as that 
from which the turning moment curve, Fig. 19, was obtained. 
This diagram, which is reproduced in Fig. 76, shows an explosion 
pressure of 262 pounds gauge per square inch, 184 pounds after 
0.1 of the stroke is completed, 138 pounds after 0.2 of the stroke 
is completed, etc. The inertia forces are obtained from Table I, 
and are subtracted from the corresponding gas pressures. The re¬ 
spective values of the gas pressures and inertia forces, and the 
resultant pressures, are given in the following table: 


Point of stroke. o.x 0.2 0.4 0.6 0.8 

Gas pressure. 184 138 88 60 44 

Inertia force. 52 35 6 —20 —36 


Resultant . 132 103 82 80 80 


The next problem is to find the values of tangent 0 for the 
different points on the stroke. The angle <P which the connecting 
rod makes with the cylinder centre line depends upon the angular 
position 0 of the crank and the ratio n of the connecting rod 
to the length of stroke— 

sin 0 = ^ (equation 18 a) 

2 n 

We also found (equation 20) that the distance of the piston head 
from the top end of the stroke is 


_ l / a\ \ 7 /sin 2 0\ 

x = — (1 — cos 0 ) nl (-I 

2 \8« 2 / 

-t(‘- 


cos« + “5l®\. 
4 n J 


Multiplying both sides by 4 n, 


4 n x — — (4 n — 4 n cos 6 -f- sin* 6 ), 


2 

Substituting 1 —cos 2 0 for sin 2 0. 










PISTON, PISTON RINGS AND PISTON PIN. 139 

t , 

4 n x -- (4 n — 4 n cos 0 + 1 — cos* 3 ). 

Dividing both sides by —and transposing, 

2 

cos 2 0 4- 4 n cos 0 = 4«-f I — 8 n — • 

l 

Completing the square, 




COS 2 0+4 n cos 0 + 4 w 2 = 4 «r 2 -j- 4 tz -f- 3: — 8 n• 

i 



Fig. 76. —Standard Indicator and Inertia Diagram. 






140 PISTON, PISTON RINGS AND PISTON PIN. 


Extracting roots, 

cos 6 -f- 2 n — 



4 n ■-)- 1 — Sn -j- • 


and, transposing, 



This equation enables us to determine directly the crank angle 
6 corresponding to any piston position. We then determine from 
a table of sines and cosines the value of the sine of 0, which we 
divide by 2 n (equation i8a.) This gives us the sine of the 
connecting rod angle 0 . Then, from a table of sines, we find the 
corresponding angle, and finally from a table of tangents, the tan¬ 
gent of this angle. The results of the various steps are given in 
the following table (for n = 2) : 


Point of stroke. o.i 0.2 0.4 0.6 

Cos (9 . 0.837 0.669 0.313 —0.076 0.507 

Sin 6 . 0.547 0.743 0.950 0.997 0.862 

Sin 0. 0.137 0.186 0.237 0.249 0.215 

0. 7° S 2 ' io° 43' 13 0 43' 14 0 25' 12 0 25' 

Tan 0. 0.138 0.189 0.244 0.257 0.220 


Multiplying the values of Pp by those of tan <t>, we get the 
following values for the piston pressure on the cylinder wall for 
different points of the stroke (in pounds oer square inch of 
piston head area) : 

Point of stroke. 0.1 0.2 0.4 0.6 0.8 

Pressure on cylinder wall.. 18.22 19.48 20.01 20.56 17.6 

In a similar manner we can determine the side pressures on the 
piston wall during the compression stroke. The gas pressure at 
different distances from the top of the stroke, the corresponding 
inertia forces and the resultant pressures are as follows: 


Point of stroke (from top). 0.1 

0.2 

0.4 

0.6 

0.8 

Gas pressure. 30 

17 

9 

6 

3 

Inertia force.— 52 

— 35 

— 6 

18 

36 

— 22 

— 18 

3 

26 

39 

Multiplying these pressures by 

the 

corresponding value 

of tan 

0 we get 

Point of stroke (from top) 0.1 

0.2 

0.4 

0.6 

0.8 

Pressure on cylinder wall..— 3.03 - 

-3.40 0.75 

6.68 

8.58 


Side Thrust Diagrams—The side pressures on the cylinder 
wall are plotted in the diagram Fig. 77. The pressure is zero at 
the beginning and end of the stroke, when the connecting rod is 
in line with the cylinder axis and tan <P = o. It rises and 
falls very quickly, however, and is almost constant throughout the 
stroke. Jn this case the average side pressure on the cylinder 

















PISTON, PISTON RINGS AND PISTON PIN. 141 


wall during the power stroke is about 17.5 pounds per square 
inch of piston head area. 

During the first portion of the compression stroke the resultant 
pressure on the piston is in the same direction as during the 
power stroke, but the angle 0 is then negative, and for this reason 
the pressure during the first portion of this stroke is on the op¬ 
posite side of the cylinder wall, and is plotted below the axis of 
abscissas. Toward the end of the compression stroke the inertia 
force exceeds the gas pressure, the piston then pulling on the 



Fig. 77.—Pressure of Piston on Cylinder Wall per Square 

Inch of Piston Head Area. 


connecting rod instead of being pushed by it, and in consequence 
the side pressure on the cylinder wall changes in direction In 
the present case the average value of the side pressure during the 
compression stroke is 2.7 pounds per square inch of piston head 
area. The side pressures during the compression stroke on the 
same side of the cylinder wall on which the pressure comes dur¬ 
ing the power stroke are considered negative, and their mean 
















142 PISTON, PISTON RINGS AND PISTON PIN. 

value is subtracted from the mean value of the positive pressures. 

The curve Fig. 77 gives the side pressure on the cylinder wall 
for one particular indicator diagram and one particular ratio of 
connecting rod length to length of stroke. This pressure naturally 
increases with an increase in explosion pressure, and it also in¬ 
creases with a decrease in the length of the connecting rod, the 
latter because the shorter the connecting rod the greater will be 
its angularity for any particular position of the piston. Chart II 
herewith gives the mean pressures on the cylinder wall during 
the power stroke per square inch of piston head area for ratios of 
connecting rod length to length of stroke varying from 1.75 to 2.5, 
and for explosion pressures varying from 225 pounds to 350 
pounds gauge per square inch. It is, of course, not only the pres¬ 
sure at the moment of explosion that determines the side pressure 
on the cylinder wall, but the pressure throughout the power stroke, 
and it should be pointed out that in calculating Chart II it was 
assumed that the gas pressure followed the law 

fiv 1 ’* — constant, 

and that the gas was expanded to 32 pounds gauge per square 
inch in every case. That the mean pressure against the cylinder 
wall found in Fig. 77 does not absolutely agree with the value 
found from Chart II is due to the fact that the calculations for 
Fig. 77 are based on an actual rather than an ideal indicator dia¬ 
gram. The discrepancy is slight, however. 

It will be seen that the side thrust does not increase in the 
same proportion as the explosion pressure. This is due to the 
fact that the pressure in the diagrams varies only at the beginning 
cf the power stroke, and is always the same at the end of the 
stroke. On the other hand, the length of the connecting rod has 
a very pronounced effect on the mean pressure on the cylinder 
wall. With a connecting rod equal to 1.75 times the length of the 
stroke the mean side pressure on the cylinder wall is roughly 50 
per cent, greater than with a connecting rod equal to 2.5 times 
the length of the stroke. ** 

'Piston Bearing Pressure—When the compression ratio and 
the ratio of connecting rod length to length of stroke have been 
determined, the mean side pressure of the piston on the cylinder 
wall can be found from Chart II. The explosion pressures con¬ 
sidered in the construction of that chart are the normal explosion 
pressures, and may be taken as three-fourths of the values given 
in Table IV. Let us assume a 4x5 inch engine with a compres¬ 
sion ratio of 4.2 and a ratio of connecting rod length to 
length of stroke of 2.1. Then the normal explosion pressure 
will be 


PISTON, PISTON RINGS AND PISTON PIN. 143 



Chart II.—Mean Pressure on Cylinder Wall During Power 
Stroke per Square Inch of Piston Head Area for Dif¬ 
ferent Relative Connecting Rod Lengths and 
Different Explosion Pressures—The Curves 
Represent Normal Explosion Pressures. 


54x368 = 276 pounds per square inch, 

and the mean pressure on the cylinder wall per square inch of 
piston head area is found from Chart II to be 16.5 pounds. The 
total pressure on the cylinder wall is, therefore, 

4X4X3. 1416 ^ — 207.2 founds. 

4 




































144 PISTON, PISTON RINGS AND PISTON PIN. 


A mean pressure of 18 pounds per square inch on the cylinder 
wall is permissible. Consequently, the projected bearing surface 
of the piston should be 

207.2 -r 18= 11.5 square inches. 

Since the piston is 4 inches in diameter this means an effective 
bearing length of 

11.5 -r 4 = 2.875 inches. 


If we figure on four rings each b/20 in width, the length occu¬ 
pied by the rings will be 4X4/20 = 0.8 inch. 

It is customary to undercut the piston for a certain length on 
each side of the piston pin axis, the chief reason for this being 
the fact that the flexure of the piston pin when under load 
is likely to deform the piston surface in the vicinity of the 
piston bosses, so that it will not bear evenly. The length of 
this under-cut portion may be made equal to 0.3 b. In the present 
case this length would be 1.2 inches. The total length of the 
piston would therefore be 

2.875 t 0.8 + 1.2 = 4.875 inches. 

The mean bearing pressure of 18 pounds per square inch 

is a good average figure. This 
pressure can be varied somewhat 
according to the purpose for 
which the motor is to be used. 
For a light motor for a speedster, 
etc., a greater pressure can be al¬ 
lowed, while for a truck motor a 
lower pressure will be advisable 
on account of the increased life 
of the piston and cylinder which 
it insures. 

Effect of Piston Friction— 

when the piston is at rest it is 
held in equilibrium by three forces, 
viz., the gas pressure P p on the 
piston head, the connecting rod 
thrust P c and the reaction of the 
cylinder wall P s against the pis¬ 
ton (Fig. 78). When, however, 
the piston is moving another 
force comes into play, namely, 
the frictional force F at the con¬ 
tact surface between the piston 
Fig. 78.— Forces on Piston an ^ cylinder. This force is equal 
in Motion. to the product of the side pres- 














PISTON, PIJ 5 TON RINGS AND PISTON PIN. 145 


sure P 8 on the cylinder wall by the friction coefficient p. The 
latter may be taken at 0.15. The direction of the force F is oppo¬ 
site to the direction of motion. The force F does not add to or 
subtract from the mean pressure against the cylinder wall, but it 
affects the distribution of that pressure. It has a tendency to 
turn the piston around the piston pin axis, the turning moment 
being FyPb/2. This turning moment is equal to a couple produced 
by two forces F/2 acting on the piston perpendicularly to the di¬ 


rection of the force F, at distances 


b 

2 


from the piston pin axis. 


Distribution of Side Thrust—The component of the con¬ 
necting rod reaction which results in side pressure on the cyl¬ 
inder wall, passes through the piston pin axis perpendicular to 
the axis of the cylinder, and is, therefore, a concentrated force. 

The reaction of the cylinder wall 
on the piston, on the other 

hand, is distributed over the en¬ 
tire bearing surface of the piston. 
If the piston had a bearing sur¬ 
face over its entire length and 
the piston pin were exactly at 
the middle of its length, then the 
reaction would be evenly dis¬ 
tributed over the entire bearing 
surface. But if the piston pin 
axis were, under these condi¬ 
tions, nearer the head end of 
the piston, for instance, then 

the reaction would be greater 
at the head end than at the 
open end of the piston. The 
specific pressures at the two 
ends under these conditions 
would be inversely proportional to their distances from 
the piston pin axis, and the pressure would decrease uni¬ 
formly from the head end to the open end. This effect is 

illustrated in Fig. 79. The grooves disturb the distribution of 

the pressure somewhat, and their presence must be taken into 
account. 



Fig. 79. —Distribution of 
Surface Pressure with 
Unsymmetrical Load. 


Best Location of Pin—What we want is a uniform distri¬ 
bution of pressure over the entire bearing length of the piston. 
It has already been pointed out that by moving the piston pin 
farther toward one end of the piston the pressure at that end 










146 PISTON, PISTON RINGS AND P 4 STON PIN. 


is increased. If there were no ring grooves the pressure would 
be equally distributed when the piston is at rest, if the piston 
pin were exactly in the middle of the piston. The presence of 
the grooves would lead us to move the pin down from the 
centre, while the couple due to the frictional force when the 
piston makes the power stroke would lead us to move it upward, 
in order to neutralize the effect of the couple. Where, then, 
is the best position for the pin? 

Let us assume that the piston in regular operation presses 

against the cylinder wall with a uniform pressure a per inch 
of length—the result we desire to attain. Let us call the length 



Fig. 8 o.—Diagram of Piston Moments. 
of the piston l, and let the piston pin axis be at a distance x 
from the head end. The total pressure of the piston against the 
cylinder wall is then 



- being the length occupied by the rings. We will assume that 
5 

the rings are evenly distributed between the head end of the pis¬ 
ton and the piston pin axis, which is generally very nearly 
the case. 

Taking moments around the piston pin axis (Fig. 80), 























PISTON, PISTON RINGS AND PISTON PIN. 147 


a (* - T)(T) +o -°M'-!-) (I) 

= a(/—x)( / —^)-°.c, 7 5 a (/-A) (A). 

Simplifying and transposing, 

Dividing by — and multiplying by 5, 

2 

5 * 2 — & * + 0.75 b l — o. 15 b 1 — 5 V- -f- 5 x 2 — 10/ x. 
Transposing, 


and 


10 lx — bx = 5l 2 -\-o.i$b 7 — 0.75 b l. 


, = 5 g .+ o..5»»- _ojS>/ .< 5 o) 

10 7 — o 

In the example of a 4x5 inch engine, which we considered 
a little further back, we found that the piston had to be 4.875 
inches long. Substituting the values 


b = 4 7=4.875. 

in equation (50) we get 

x = (5 X 4I X 4 s) + (0.15X4X4) — (o -75 X 4 X 4I) _ 2 g 

(10 X 4i) — 4 

inches. 

That is the piston pin axis should be 2.38 inches from the 
head end, which is one sixteenth of an inch above the centre 
of the piston. 

Thickness of Piston Walls—It is almost the universal prac¬ 
tice to connect the piston head and piston bosses by a rib of 
substantially the form shown in Fig. 81. This rib considerably 
strengthens the cylinder head and permits it to be made thinner 
than if no rib were provided. It also relieves the side wall of 
the piston of some strain and permits of this wall being made 
thinner. Pistons of very small diameter are occasionally made 
without this rib, while pistons of large size sometimes have 
four or six equally spaced ribs joining the head to the side wall 
of the piston. 

The head can also be strengthened by making it convex or 
bulging outwardly, and this form of head is frequently met 
with, a second reason for its use being that when a high com¬ 
pression ratio is desired it helps to reduce the compression vol¬ 
ume. It is, however, a better plan to make the head plane, as in 
that case its surface in contact with the hot gases is a minimum, 
and the piston will then absorb less heat. Since it is not provided 








148 PISTON, PISTON RINGS AND PISTON PIN. 

with any cooling means this is a matter of considerable moment. 

In developing formulae for the thickness of the different 
parts of the piston wall, it must be borne in mind that all of 
these walls are machined on one side only, and some allowance 
must be made for inaccurate core work. That is to say, the for¬ 
mulae for the thicknesses should be of the form 

t = a b -f c, 

where t is the thickness of the wall, b the bore and a and c are 
constants, c being allowance for inaccuracies in the casting. 
In pistons of the type represented in Fig. 81, the head thickness 
may be made 

/h = 0.032 b -j- 0.060 inch .(51) 

If eccentric rings are used the maximum thickness of f:he ring is 


-- — 0.0364 b, 

27-5 

and the depth of the groove must be made about 10 per cent, 
greater, viz., 

c? g = o.04&..(52) 

The thickness of the piston wall where the ring grooves are 
located can be determined by the equation 

tu = 0.062 b -f- o. 10 inch .(53) 

and the thickness at the lower end by the equation 

/] = 0.02 b 0.05 inch .(54) 


In using equations (51) to (54) the nearest dimension in thirty- 
seconds of an inch can be chosen. 

Allowance for Heat Expansion—The average American 
practice in regard to piston clearance allowance is to allow 2 
thousandths of an inch per inch of cylinder bore at the head 
end of the piston, and $4 thousandth at the bottom end, the piston 
being tapered from the head end toward the bottom end. The 
piston diameter at the head end, therefore, should be 


d b =0.998 b .(55) 

and the piston diameter at the bottom end, 

db = 0.9995 b ..(56) 


Some makers reduce only the uppermost land and make the 
remainder of the piston cylindrical. 

Number and Location of Rings.—Practice is about evenly 
divided between the use of three and of four rings. If a light, 
compact motor is desired, it will be advisable to limit the num¬ 
ber of rings to three, while if lightness is not such an important 
consideration, it will be better to use four. Generally, all of the 
rings are placed on that part of the piston above the piston pin, 
but occasionally a ring is placed near the lower or open end. 
This is an inheritance from stationary gas engine practice. It 









PISTON, PISTON RINGS AND PISTON PIN. 149 


is sometimes claimed that the object of this ring is to aid in the 
distribution of the oil, but the original purpose for which the 
ring was placed at the lower end was to have it over-travel the 
bottom end of the cylinder bore, and thus prevent the forming 
of a ridge where the travel of the lowest ring ends. It is usual 
to cut two or three circular grooves in the outer surfaces of the 
piston near the lower end for the distribution of the oil. 

Sample Design—Having thus derived formulae for all the 
principal piston dimensions, we are now in position to design a 
piston for our 4x5 inch motor. The required length has already 
been determined to be 4.875 inches, and the best location of the 



piston pin axis as 2.38 (2^4) inches from the head end The 
diameter of the piston at the head end will be 

4X0.998=3.992 inches, 
and the diameter at the bottom end 

4X0.9995=3.998 inches. 

The thickness of the head— 

4X0.032-1-0.060=0.188 (3/16 inch). 

The depth of the ring grooves— 

4X0.04=0.160 (5/32 inch). 

The thickness of the piston wall at the upper end— 
4X0.062+0.10=0.348 (11/32 inch), 








































































150 PISTON, PISTON RINGS AND PISTON PIN. 


and the thickness of the piston wall at the lower end— 

4X0.02 + 0.05=0.13 (1/8 inch). 

The necessary diameter of the piston pin has already been 
found to be 1 inch, that of the piston bosses 1 V2. inches, and 
the length of the piston bosses, 1 inch. The length of the under¬ 
cut portion of the piston, which is symmetrical with the piston pin 
axis, will be 

4X0.3 = 1.2 (1 3/16 inches). 

This undercut may be made 1/64 of an inch; that is, the piston 
at this portion may be reduced to 3 31/32 inches in diameter. 
The width of the piston rings is 

44-20 = 0.200 (7/32 inch.) 

It will be found that four rings can be placed on the top end of 
the piston and each of the “lands” between rings made 3/16 inch 
wide. 

Since the wall of the piston at the lower end is usually made 
very thin, it is a good plan to provide the casting with an in¬ 
ternal flange at the lower end to make it more rigid, which 
is especially desirable for machining. 

Manufacture of Pistons—Since the walls of piston cast¬ 
ings are very thin, they are often hard in spots, and for this 
reason the castings are first annealed. They are filled with 
charcoal and brought to a good red heat in an annealing furnace, 
and are then allowed to cool slowly on a bed of ground bone 
so that the heads of the pistons will not chill. After annealing 
the castings are sand blasted to remove the core sand, and if it 
is thought desirable to clean them still further, they may be 
placed in a tank of kerosene. 

The first machining operations on a piston of the type illus¬ 
trated in Fig. 81 consist in centre drilling and reaming the 
little boss generally provided on the head, and boring out the 
internal flange or rib at the open end. The latter operation can 
be accomplished to advantage by means of a boring tool and 
floating reamer in a turret lathe. 

The next operations consist of turning the piston on the out¬ 
side, facing the head and cutting the ring grooves. These opera¬ 
tions are preferably performed at one setting in a turret lathe 
specially adapted for the work. A face plate for holding and 
driving the piston and the necessary tools for finishing it, are 
illustrated in Fig. 82. The piston is secured to the face plate A 
by means of the draw-in rod B, with a Tee head which engages 
/ver the piston bosses. It is centred and held square to the 
face plate by the finished internal flange at its open end. The 


PISTON, PISTON RINGS AND PISTON PIN. 151 



Fig. 82.—Turning and Grooving Piston in Turret Lathe. 


casting is driven by means of a semi-floating driver E attached 
to the face plate A. This driver bears uniformly on the two 
bosses and floats on the flat key F at right angles to the driving 
faces. The casting is further steadied by means of the tail 



Fig. 83.—Arbor for Grinding Piston. 












































































































































152 PISTON, PISTON RINGS AND PISTON PIN. 

centre. 1 he turret head on the cross slide contains the tools 
for rough turning, rough facing, rough grooving and under cut¬ 
ting, which are brought into operation in the order given. Finish 
facing, finish grooving and cutting of oil grooves are accom¬ 
plished simultaneously by means of a set of tools carried on the 
back end of the carriage. These turret lathes are provided 
with star wheel stops and with spring actuated ball locating 
pawls on the head stock, which makes longitudinal measurements 
unnecessary. If the piston is of such design that the rib prevents 
the T head of the draw-in rod from engaging securely over the 
piston bosses, the bosses may first be drilled and a slip pin used 
instead of the T head. 






'///// 

v 22 Zl 



Fig. 84.—Piston Boss Drilling Jig. 







































































































PISTON, PISTON RINGS AND PISTON PIN. 153 




Fig. 85.—Jig for Drilling Lock Screw Hole. 

If the piston is to be ground on the outside, no finishing 
cut is necessary, but an allowance for grinding of from 0.020 
to 0.030 is made in the diameter. An arbor suitable for 
holding the piston for grinding is illustrated in Fig. 83. The 
grinding is done after the piston bosses have been drilled and 
reamed, and the sliding rod A is passed through the holes in the 
piston bosses, while the sliding collar B is drawn up against the 
finished open end of the piston by means of the nut C. The 
piston being centre drilled and reamed at the head end, when it 
is secured on the arbor, the assembly can be supported in the 
grinding machine between centres. 

For drilling and reaming the holes in the piston bosses a 
drilling fixture is made use of, which can be secured to a lathe 
face plate or screwed to the head stock. A well designed jig 
for this purpose, which was described by Robert G. Pilkington 
















































154 PISTON, PISTON RINGS AND PISTON PIN. 

in The Horseless Age of October 21, 1908, is illustrated in 
Fig. 84. This jig is adapted to be screwed to the head stock 
of a turret lathe. It consists essentially of a cylindrical casting 
into which the piston casting fits, grooves being left on opposite 
sides for the escape of the chips. One head of the jig is perma¬ 
nently fastened in place and is provided with lugs which straddle 
the piston bosses so as to centre them with the tool. The other 
head is in the form of a hinged yoke which can be swung aside 
to introduce the piston into the fixture and is held in place by 
means of a swinging stirrup and clamp screw. 

On the entering side of the jig slip bushings are used, of 
the proper size for the drill and reamer, respectively. The 
drill is guided only by the bushing on the entering side, while 
the reamer is guided at both ends. 

In drilling the holes for the piston pin lock screw the jig 
shown in Fig. 85, also due to Pilkington, can be used to advan¬ 
tage. This jig is made in two parts, a body and a clamp, which 
form handles convenient to grasp in one hand while the other 
feeds the drill. When so held the jig is firmly clamped to the 
piston and will not be pulled off when the drill is raised out 
of the bushing. A hardened steel guide pin is placed in the 
piston pin hole, so as to centre the bushings with the hole by 
means of machined surfaces straddling the guide pin. 




CHAPTER VIII. 


OFFSET CYLINDERS. 

Object of Offsetting—The side thrust of the piston against 
the cylinder wall is the cause of both wear of the piston and 
cylinder and of frictional loss, and any means of reducing it is 
therefore doubly advantageous. It was shown in a previous 
chapter that this side thrust is proportional to the effective 
pressure on the piston and to the tangent of the connecting rod 
angle. The effective pressure on the piston cannot well be 
reduced, as upon it depends the power developed by the motor, 
so the only chance lies in reducing the value of tangent <t> (that 
is, of the angle itself). The most obvious way of accomplish¬ 
ing this is by making the connecting rod longer. Old text books 
on the steam engine dwell at considerable length on the ideal 
engine with a connecting rod of infinite length, in which, among 
other things, there is no side thrust on the cylinder wall. In 
practice, however, the length of the connecting rod is limited, 
since the height of the motor, and consequently its bulk and 
weight, increase rapidly with this length. There is, however, a 
method of arranging the moving parts so that with a fixed length 
of connecting rod the connecting rod angle will be relatively 
small when the pressure on the piston is high, and proportion¬ 
ately greater when the pressure on the piston is low. This con¬ 
sists in offsetting the cylinder axis from the vertical plane 
through the crank shaft axis toward the side on which the crank 
arm moves during the power stroke. The angularity of the 
connecting rod for any position on the power stroke, except at 
the very ends of the stroke, is then less than it is for the corre¬ 
sponding position in a motor without offset cylinders—a sym¬ 
metrical motor. 

Offsetting of the cylinders (or of the crank shaft) has some 
rather peculiar effects on the functions of the motor. In the 
first place, the piston is not at the ends of its stroke when the 
crank arm is in the vertical positions; it reaches the ends of the 

155 



156 


OFFSET CYLINDERS. 


stroke a moment later, when the crank arm and connecting rod 
are in line with each other. Besides, the two dead centre posi¬ 
tions are not 180 degrees apart, as they are in a symmetrical 
motor, but slightly less, as clearly shown in Fig. 86. The 
effective or actual length of the stroke, instead of being equal 
to twice the length of the crank arm, is slightly greater than 
this. The connecting rod is parallel with the cylinder axis 




tions in Offset Cylinder Cylinder. 

Motor. 

some time after the top vertical position of the crank has been 
passed and the same time before the bottom vertical position is 
reached. The two points for which the connecting rod angle 
becomes zero both fall in the downward stroke. The piston is 
not at the same distance from the top end, say, of the stroke, 
when the crank arm is at the same angular distance from its 




































OFFSET CYLINDERS. 157 


upper vertical position on opposite sides of the centre, re¬ 
spectively, as in the case in the symmetrical motor. 

Length of Stroke and Dead Centre Positions—Referring 
to Fig. 86, when the crank and connecting rod are in the upper 
dead centre position, the sine of the crank angle 

o = _ L _ f 


sm 


L + „ / 1 (« + i) 
2 


( 57 ) 


and when the piston is in the bottom dead centre position 
sin Q x = — - t- - — — f 


n l — — 
2 


l (n — 1) 


( 53 ) 


In our particular example, with an offset of — 1 inch. 

4 


sin 0 — 


and 

also 


5 X 2.5 

*= 4 ° 35': 


— = 0.08, 


and 


sin 6 , = — l -= — o. 133 

5 X 1.5 


0 X = 18 7 0 40 7 . 

If we assume that the crank motion is uniform, which is very 
nearly the case, it follows that the mean piston speed is slightly 
less during the down stroke than during the up stroke, since 
the crank travels through a greater angular distance during the 
down stroke than during the up stroke. 

Referring to Fig. 86, it can easily be seen that the actual length 
of stroke is 


5= j/ («/+ -f) /(»'- ~) '->= 

|/>(« + D 2 -/*--|/ /•(»--! y-r . (59) 

For a motor of 5 inch nominal stroke with 10 inch connecting 
rod and 1 inch offset this gives 

s - 5.027 inches, 

while for the same motor with an offset of 0.666 inch (one- 
sixth the bore) it gives 

5--5.012 inches. 

The main object of offsetting the cylinder being a reduction in 
the side thrust of the piston against' the cylinder wall, we will 
investigate the variation of this side thrust in an offset cylinder 














158 


OFFSET CYLINDERS. 


motor during the power stroke and compression stroke, respec¬ 
tively. We will also investigate the effect of offsetting on the 
timing of the motor and on the balance of the reciprocating 
parts. To this end it is necessary to develop equations for the 
piston position, piston speed and piston acceleration correspond¬ 
ing to any crank angle, the same as was done for the symmetrical 
motor. 

Piston Speed and Acceleration.—Referring to Fig. 87, let f 
represent the amount of the offset. Let the length of the con¬ 
necting rod be denoted by l n and the length ot the crank arm by 


— , (though it should be pointed out that in this case l stands for 
2 

the nominal length of stroke and not the actual). The use of the 
same symbols as in the previous discussion has the advantage 
that the equations developed should reduce to the corresponding 
equations for the symmetrical motor if f is made o, which will 
be a check on the accuracy of the work. The crank and con¬ 
necting rod angles are again represented by 0 and 0, respec¬ 
tively. As regards the momentary piston position, it is more 
convenient in this case to represent this by the vertical distance 
between the piston pin axis and the crank shaft axis, rather 
than by the distance between the momentary position of the pis¬ 
ton pin axis and its highest position. 

Let x represent the vertical distance between the piston pin 
axis and the crank shaft axis. Then (Fig. 87), 


and 


x = n l cos 0 -f- — cos 6 , 
2 

n /sin 0 -f-/ = — sin 6 . 

2 


This latter relation holds for any point of the crank circle. 
It follows that 


sin 0 — 


— sin 6 —f 
2 


n l 


sin 6 f 

2 n n l 


Hence, 


But cos 0 = •J 1 — sin 2 0 = 



sin 0 
2 n 



(60) 



In order to get rid of the radical expression we add to it 










OFFSET CYLINDERS 


159 


i 

i 



which makes it a perfect square and does not introduce an 
appreciable error. Extracting the root, we get 


7 nl /sin 0 f \ 2 l a 
nl — — (--— ) i —cos 6 

2 \ 2 n n 1} ‘2 


,(62) 


If now we differentiate this expression with respect to time 
we get the momentary piston speed 


But 

hence, 


v — 


d_ 

d 


t L \ 2 n n ij 2 n 2 

d 6 7r N , , 

- radians fer second , 

dt~ 30 


d 0 
d t 


v— — /— sin 0 cos 0 — cos 0 -f- — sin • 

30 \4 n 2 n 2 ) 

Dividing by 12 to reduce to feet per second, and multiplying 
the expression in parentheses by 2, and dividing the coefficient 
of this expression by the same number, 

v = — ?r ^ r sin 0 cos 0 — — cos 0 + / sin 0 \ feet f>er second. .(63) 
720 \2 n n ) 

When f = o this equation becomes identical with equation (21), 
giving the momentary piston speed for a symmetrical motor. 

In order to find the piston acceleration we differentiate equa¬ 
tion (63), which gives 


d v = — _Z_ (cos 2 0 — sin 2 0) -f- sin 0 -|- / cos 0"] d 0 

720 \_2 n n 

and remembering that 


and that 
we get 


d 0 7T JV ^ 
d t~ 30 

cos 2 0 — sin 2 0 = cos 2 0 , 


dv tt 2 JV 2 (l a \ f • a \ 1 A 

a — — — —-( — cos 2 0 -+- — sin 0 + / cos 0 ) = 

d t 21,600 \2 n n ] 


— (JL cos 2 0 + sin 0 + cos .(6 4 > 

21,600 \2 n n L / 

This equation becomes identical with equation (23) (which 
gives the piston acceleration for a symmetrical motor) when the 
offset f is zero, except that it is preceded by the minus sign, 
which is due to the fact that in this case we are considering the 
motion of the piston pin axis relative to the crank shaft axis, 
which is opposite to the motion of the piston pin relative to its 
highest position, which we considered in the former case. 












160 


OFFSET CYLINDERS. 


If W represents the weight of the reciprocating parts, then the 
inertia force in the offset motor may be represented by the equa¬ 
tion 

F* = 0.0000142 Wl(— — cos 2 0 4“ — sin 0 4“ cos .(65) 

\2 n n l J 

Side Thrust on Cylinder Wall—The length of the offset 
employed in practice lies generally within the limits of one- 
quarter the bore and one-sixth the bore. American practice 
rather favors an offset of one-sixth, while European practice 
favors an offset of one-fourth. In determining the side thrust 
for different points of the up stroke and the down stroke we 
must first determine the crank angle corresponding to different 
points on both strokes. The deduction of an equation giving this 



Fig. 88 .—Crank Angle-Stroke Curve for Motor With an 
Offset Equal to One-fifth the Stroke. 

angle or a function of it directly involves considerable difficulty, 
and the best method is to proceed backwards, assuming different 
crank angles and then determining the corresponding point on 
the stroke by means of equation (61). From the results thus 
obtained a curve like that shown in Fig. 88 can be drawn. It is, 
of course, also possible to determine the points for this curve 
graphically without calculation, by laying off the connecting rod 
and crank arm for different angular positions of the latter. 
From Fig. 88 we can readily determine the crank angle corre¬ 
sponding to any point on the stroke. The connecting rod angle 
is next determined by means of equation (60) and the effective 




OFFSET CYLINDERS. 


161 


piston pressure from the standard diagram Fig. 76. Multiply¬ 
ing corresponding values of the connecting rod angle and the 
effective piston pressure together, we obtain the side thrust on 
the cylinder wall. 

In this way the curves in Fig. 89 were plotted. The calcula¬ 
tions were based on the assumption of a cylinder of 4 inch bore 
by 5 inch stroke. The full line curve gives the side thrust of 
the piston on the cylinder wall per square inch of piston head 
area for an offset equal to one-sixth the bore, and the dotted 
line the corresponding values for an offset equal to one-quarter 
the bore. In a 4x5 inch motor one-sixth of the bore corresponds 
to two-fifteenths of the stroke and one-quarter of the bore to 
one-fifth the stroke. In the figure the offsets are given in terms 
of the stroke so as to make the diagram applicable to all sizes 
of motors. For the sake of comparison, the side thrust diagram 
for a symmetrical motor which was given in Fig. 77 is also 
drawn in in dashed lines. Since we are not so much concerned j 
with what side of the cylinder the piston presses against, as with > 
the total pressure with which it bears against the cylinder walls, 
all side thrusts are plotted on the same side of the base line, 
irrespective of their direction. 

By reference to Fig. 87 it can readily be seen that in an offset 
motor the connecting rod angle is comparatively small through¬ 
out the power stroke, and comparatively large throughout the 
compression stroke. For this reason the side thrust on the cylin¬ 
der wall in an offset motor is less than in a symmetrical motor 
during the power stroke and greater than in a symmetrical motor 
during the compression stroke. This is clearly shown in Fig. 89. 

In a symmetrical motor the side thrust becomes zero when the 
crank is in the two vertical positions. In an offset motor, on ; 
the other hand, when the crank is in its lowest position (repre¬ 
sented in the diagram by the left hand co-ordinate axis), 
the connecting rod makes an appreciable angle with the cylinder 
axis. During the first part of the compression stroke, the greater 
the offset the greater the side thrust, because the greater the 
angularity of the connecting rod. The side thrust becomes zero 
in all three types of motors when the piston is slightly less than 
four-tenths of the stroke from its top position, because, in the 
standard diagram Fig. 76 the inertia and gas pressure then 
neutralize each other so that the effective piston pressure is zero. 

At the end of the compression stroke the side thrust is again 
zero, because at this point of the diagram the inertia and com¬ 
pression pressure again balance each other. From this point on 
the side thrust in the symmetrical motor increases rapidly; in the 





162 OFFSET CYLINDERS 



Fig. 8q.—Side Thrust Diagram. 



















OFFSET CYLINDERS. 


163 


offset cylinder motor, on the other hand, it increases for a moment 
and then drops back to zero again, which it attains when the 
connecting rod centre coincides with the cylinder axis; from that 
point on the side thrust increases in about the same manner as 
in the symmetrical motor, but decreases to zero again a moment 
before the crank reaches the lowest position, when the connecting 
rod is again in line with the cylinder axis. 

The criterion of the value of offsetting is the proportion in 
which it reduces the mean side thrust on the cylinder wall dur¬ 
ing the compression and power strokes. It will be seen from 
Fig. 89 that under the conditions here assumed, the mean side 
thrust on the cylinder wall during the compression and power 
strokes is 11.2 pounds per square inch of piston head area in a 
symmetrical motor; 8.9 pounds in a motor with an offset equal 
to two-fifteenths of the stroke, and 7.8 pounds in a motor with 
an offset equal to one-fifth the stroke. Offsetting the cylinder 
by an amount equal to two-fifteenths the stroke results therefore 
in a reduction of 20.5 per cent, in the side thrust on the cylinder 
wall during the compression and power strokes, and offsetting 
by an amount equal to one-fifth the stroke in a reduction of about 
30 per cent. 

In order to give an idea of what this reduction in the side 
thrust amounts to from the power standpoint, let us assume a 
four cylinder, 4x5 inch motor running at a piston speed of 1,000 
feet per minute. The travel during the compression and power 
strokes will correspond to 500 feet per minute. With a cylinder 
offset of one inch, equal to one-fifth the stroke, the side thrust 
will be reduced 3.4 pounds per square inch of piston head area. 
If we assume that the friction coefficient is 0.15, the friction 
between the piston and the cylinder wall will be reduced 

0.15x3.4 = 0.51 pound 

per square inch of piston head area. The four pistons together 
have an area of about 50 square inches, so the total reduction 
of the friction is 

50x0.51=25.5 pounds. 

Multiplying this by the piston travel, we get the reduction in 
the frictional loss, viz.: 

500x25.5 = 12,750 foot-pounds per minute = 0.386 horse power. 

A motor of this size would develop about 25 horse power, hence 
the gain in output would be only about 1.5 per cent. This may 
seem quite small, but when it is considered that the power thus 
saved is otherwise expended largely in wearing out the piston 
and cylinder, the matter appears in a somewhat different light. 


OFFSET CYLINDERS. 


IG4 

Effect on Timing—The valves should begin to open and 
should close and the spark in a motor should occur when the 
piston is at particular points of its stroke, and since any particu¬ 
lar point of the stroke corresponds to a different crank angle in 
an offset motor than in a symmetrical motor, it is plain that these 
functions should occur at different angular positions of the crank 
in an offset motor than in a symmetrical rr itor. If the designer 
knows of a valve timing system which gives satisfaction in a sym¬ 
metrical motor and wants to apply the same system in an offset 
motor, he should first determine, by means of equation (19), the 
piston positions corresponding to the different crank angles used 
in the timing system in the symmetrical motor, and then deter- 



Fig. 90.—Inertia Force Diagram for a Cylinder With an 
Offset of One-fitth the Stroke. 


mine by means of Fig. 88 (or a similar curve if the offset is 
more or less than one-fifth the stroke) the crank angles corre¬ 
sponding to these piston positions in an offset motor. 

Effect on Balance—An inertia force curve for a 4x5 inch 
motor with an offset of one inch, reciprocating parts weighing 
0.55 pound per square inch of piston head area and operating at 
1,000 feet piston speed per minute, is shown in Fig. 90. The data 
from which this curve was drawn were calculated by means of 
equation (65). It will be noticed that the curve is considerably 
distorted. When the crank is in the vertical positions the inertia 
forces are exactly the same as in a corresponding symmetrical 
motor, as may be seen by a comparison with Fig. 15. The inertia 
reaches both of its maxima during the upward stroke. 

In Fig. 91 one-half of this inertia curve is shifted 180 degrees 













OFFSET CYLINDERS. 


165 


with relation to the other. This represents the conditions in a 
four cylinder motor in which one pair of crank pins begins to 
move downward while the other pair begins to move upward, the 
motion of one pair being advanced 180 degrees with relation to 
that of the other. The two portions of the inertia curve are 
drawn in in dotted lirps and their resultant is drawn in in a full 
line. This resultant is exactly the same as was found in the case 
of a corresponding symmetrical motor (Fig. 30) ; consequently, 
offsetting of the cylinder has absolutely no effect on the unbal- 



Fig. 91.—Inertia Force Diagram for a Four-Cylinder Motor 
With an Offset Equal to One-fifth the Stroke. 

anced inertia force in a motor. That this is so can also easily be 
proven analytically. 

In a four cylinder motor there are two sets of moving masses 
separated by 180 degrees of crank motion. The inertia of one set 
may be represented by the equation 

Fa = 0.0000142 IVl A 72 ( •— cos 2 6 -f- sin 0 + — cos 6 \ 

\ 2 n n l 2 / 









166 OFFSET CYLINDERS, 

and that ot the other by 

Fa\ = 0.0000142 IV lN 2 cos 2 {6 -j- 180°) -|-sin (0 + i8o°)-f- 

(_2 n n l 

L cos (0 -}- 180 0 ) 

2 

The resultant unbalanced inertia force is the sum of these two 
expressions. 

But 

sin 6 = — sin (6 -f- 180°) 

and 

cos 6 = — cos (6 -f- 180°), 

hence the terms involving the factors sin 0 and cos 0 cancel out 
in the sum. On the other hand, 

cos 2 (0 + 180 0 ) = cos (2 0 + 360°) = cos 2 0 . 

Hence 

Fa, -f- Fai — O.OOOOI42 IVIN 1 ( -i-COS 2 

\ » 

This is the equation of the unbalanced inertia force. It will be 
seen that it does not involve the factor f, and the same result 
is obtained if the expression for the unbalanced inertia force in 
? four cylinder symmetrical motor is made the basis of the 
discussion. 

The reduction in the side thrust due to the use of offset cylin¬ 
ders can be taken advantage of in different ways. Either the 
purchaser may be given the benefit of a motor with reduced 
specific piston pressure, and consequent longer life, or the manu¬ 
facturer may discount these advantages, so to speak, by reducing 
the length of the piston or of the connecting rod so as to keep 
the side thrust per square inch of piston bearing surface constant. 

For instance, in the example of a piston for a 4x5 inch motor, 
considered in the previous chapter, we found that an effective 
bearing length of 2.875 inches was required. If the motor had an 
offset of 1 inch, the mean side thrust of the piston would be 
reduced 30 per cent., and the effective bearing length could there- 
fore be reduced by 30 per cent, of 2.875 inches, which is 0.86 inch. 
That is, the piston would have to be only 4 inches long instead 
of 4inches, for the same mean side thrust. 

With motors of comparatively long stroke offsetting of the 
cylinders involves some difficulty, because there usually is not 
sufficient clearance for the connecting rod during the power 
stroke. This difficulty can be overcome by casting the bottom 
end of the cylinder with a sort of slot, into which the connecting 
rod can swing, or curving the connecting rod near its lower end. 




CHAPTER IX. 


THE CRANKSHAFT. 


One of the most important parts of the gasoline motor is the 
crankshaft. It is an expensive part, requiring a very considerable 
amount of accurate machine work to be done upon it, and it is 
subjected to very severe strains, consequently it is liable to fail 
in service unless properly dimensioned and made of a suitable 
grade of material properly treated. 

Crankshaft Material—Medium carbon steel is the material 
most extensively used for crankshafts. The blanks from which 
the crankshafts are machined are generally produced by the drop 
forging process, though where a high grade of car is turned out 
in comparatively small numbers the crankshafts are sometimes 
turned up out of solid billets. The drop forging process re¬ 
quires several heatings to a good red heat; this heating impairs 
the physical qualities of the steel, and the latter have to be 
restored by suitable heat treatment, consisting of annealing, 
quenching and reheating. The carbon steel generally used for 
crankshafts has approximately the following composition and 
physical properties (after heat treatment) : 


Carbon .. 
Manganese 
Silicon . .. 
Sulphur .. 
Phosphorus 


CHEMICAL COMPOSITION. 

Per Cent. 

. 0.45 

. 0.60 

.not over 0.18 

.not over 0.04 

.not over 0.04 


PHYSICAL PROPERTIES. 


Tensile strength. 

Elastic limit. 

Elongation in 2 inches 
Contraction of area... 


90,000 lbs. per sq. in. 
70,000 lbs. per sq. in. 
15 per cent. 

38 per cent. 


A suitable heat treatment for steel of this character is as 
follows: The forgings are first heated to a temperature of about 
i,5oo° Fahr. and quenched in oil. They are then reheated to 

167 












108 the crankshaft. 

1,450° Fahr. and allowed to anneal (cool slowly) in lime. They 
are heated again to i, 45 °° Fahr. and quenched in oil, and are 
then drawn to about 900° Fahr. The drawing temperatures may 
be varied more or less, a lower drawing temperature gi\ing a 
stronger shaft and a higher drawing temperature a tougher or 
more ductile shaft. 

In a number of the higher grades of automobile motors 3V2 
per cent, nickel steel is used for the crankshaft. The (approxi¬ 
mate) composition and physical properties of this steel (after 
heat treatment) are as follows: 


Carbon ... 
Nickel .... 
Manganese 
Silicon . . . 
Sulphur . . 
Phosphorus 


CHEMICAL COMPOSITION. 

Per Cent. 

. 0.25 

. 3.5 

. 0.60 

_ # not over 0.18 

.not over 0.04 

.not over 0.04 


PHYSICAL PROPERTIES. 


Tensile strength..120,000 lbs. per sq. in. 

Elastic limit. 90,000 lbs. per sq. in. 

Elongation in 2 inches. 16 per cent. 

Contraction of area. 50 per cent. 

The heat treatment for this steel is simpler than that for 
carbon steel, and consists merely in quenching in oil at a tem¬ 
perature of 1,530° Fahr., reheating to 900° Fahr., or thereabouts, 
and then cooling slowly. 

Other materials used in the construction of crankshafts are 
chrome-nickel steel and chrome-vanadium steel. Both of these 
can be given exceedingly high tensile strength combined with a 
high ductility. A chrome-vanadium steel, for instance, incor¬ 
porating 1 per cent, of chromium and from 0.16 to 0.18 per cent, 
of vanadium, has, according to J. Kent Smith, a tensile strength 
of 127,800 pounds, an elastic limit of 110,100 pounds, an elonga¬ 
tion of 20 per cent, and a contraction of area of 58 per cent., 
when quenched in lard or fish oil at 1,650° Fahr., annealed for 
two hours at 1,000° Fahr. and cooled in air. Even better proper¬ 
ties can be obtained with chrome-nickel steel when suitably 
treated, but owing to the fact that this material is very hard to 
machine, it is little used for crankshafts. Vanadium steel ma¬ 
chines with less difficulty. Vanadium when added to steel con¬ 
taining no other metallic alloys in appreciable quantity con¬ 
siderably improves the endurance properties of the steel, but in 
order to materially increase the tensile strength the addition 
of another metal is necessary, such as nickel, chromium or 
tungsten. 













THE CRANKSHAFT. 


Kill 

It may be pointed out here that the physical properties of the 
material are generally left out of account in determining the 
dimensions of the crankshaft. The reason for this is that one 
of the chief requirements in a crankshaft is stiffness or rigidity. 
This depends upon the coefficient of elasticity of the material, 
which is generally believed to be about the same for all grades 
of steel (in the neighborhood of 30,000,000), though Dr. Sargent, 
of the Carpenter Steel Company, claims that chrome-nickel steel 
when properly heat treated has a coefficient of elasticity of 50,- 
000,000. 

Types of Crankshafts—A crankshaft is composed of the 
crank pins, crank arms, crank jon*"" '3 and driving ends. As a rule, 
crankshafts are made from integral forgings, but occasionally they 

, built up. Built-up crankshafts are used for two reasons. 
In the first place, it is quite customary abroad to enclose the 
flywheels of small single and double cylinder motors in the crank 
case, and the flywheels then take the place of the crank arms, the 
crank pin and crank journals being bolted to the flywheels by 
means of nuts and tapered seats in hubs cast on the flywheels. 
These flywheels are cast with solid webs. The other reason for 
using built-up crankshafts is that if ball bearings are to be used 
on the intermediate journals, in the case of an integral crankshaft 
the inner race has to be of such a large diameter that it can be 
stripped over the crank arms. To avoid this the cranks are 
built up of a number of parts in such a manner that ball bearings 
with a bore of the inner race equal to the journal diameter can 
be put in place on the journal. 

The most common types of crankshafts for one, two, four and 
six cylinder motors are illustrated in Fig. 92. It should here be 
pointed out that a crank pin, together with the two crank arms 
on opposite sides of it, is frequently referred to as a “throw.”' 
By reference to the figure it will be seen that in some crank¬ 
shafts there is only a single throw between a pair of crank jour¬ 
nals or supporting bearings; in other crankshafts there are two 
throws between crank journals; in others three, and in still 
others four. The dimensions of the crank pins are naturally 
largely determined by their function as a bearing, and it might 
be thought that the unit pressure on the pins should be approxi¬ 
mately the same in crankshafts of all types and sizes. However, 
when there are several throws between supporting bearings the 
pins must necessarily be made of large diameter, so as to be able 
to withstand the bending stresses to which they are subjected at 
the moment of explosion, and since with a larger pin diameter 
the rubbing speed is greater, the unit pressure must be reduced 
if there is not to be undue heating at the bearing. 


170 


THE CRANKSHAFT. 



Fig. 92.—Types of Crankshafts. 






























































































































































































































THE CRANKSHAFT. 


171 


Balancing Crankshafts—In a single cylinder crankshaft the 
centrifugal force on the crank arms, crank pin and part of the 
connecting rod forms an unbalanced rotating force which would 
cause a great deal of vibration if no means were provided for 
properly balancing it. The method generally resorted to con¬ 
sists in applying balance weights to the crank arms, as pointed 
out in the chapter on Balancing of Engines. The method of 
calculating the proper balance weight was there given. The bal¬ 
ance weight should be such that the centrifugal force acting on it 
is equal to the centrifugal force acting on the crank arms, crank 
pin and half of the connecting rod, plus one-half of the maxi¬ 
mum inertia force acting on the reciprocating parts. 

The balance weights in a high speed motor must be securely 
fastened to the crank arms, since the strains due to the centrif- 



Fig. 93.— Methods of Fastening Balance Weights. 


ugai force at ‘'racing” speeds are very considerable, and the 
loosening of one of the weights would be sure to do serious 
damage. Three methods of fastening the balance weights are 
illustrated in Fig. 93. The most secure way is undoubtedly to 
extend the crank arms to the opposite side from the crank pin, 
milling a transverse groove into the arms on the outer side near 
the end, fitting a counterweight with a tongue into this groove, 
and bolting it in olace (see A, Fig. 93). The clamping bolt 
is then relieved of all shearing stresses. A method that is ex¬ 
tensively used in marine practice is represented at B in the same 
figure. In this case the counterweight is held in place by a bolt 




















































172 


THE CRANKSHAFT. 


which passes through the end of the crank arm, and through por 
tions of the counterweight passing over the end of the arm. At 
C is shown the method employed in the Brush single cylinder 
motor, the counterweight being held in place by a stud and nut, 
the stud being screwed into the crank arm from the end. In 
the Brush motor the counterweights are filled with lead to get 
the necessary weight into a small space. 

In a double cylinder opposed motor the crankshaft is always 
made with two throws set at 180 degrees relative to each other. 
In this way one set of reciprocating parts always moves opposite 
to the other set, and at exactly the same speed, so that the re¬ 
ciprocating parts are perfectly balanced, except for the fact that 
the two sets are not quite in line with each other. The rotating 
parts are also very nearly balanced, since the centrifugal force 
acting on one throw is equal and opposite in direction to the cen¬ 
trifugal force acting on the other throw, and there is only a 
small couple due to the centrifugal forces acting on arms equal 
to half the distance between the cylinder axes. This is gener¬ 
ally so slight that it can be neglected, and no special means for 
balancing such a crankshaft are needed. 

In a four cylinder vertical motor the four throws are always 
in the same plane, the two outer throws being on the same side 
of the crankshaft centre, and the two inner throws on the op¬ 
posite side. The centrifugal force acting on each throw is a 
radial rotating force which may be considered to act at the 
centre of that throw. The centrifugal forces F on the two outer 
throws (Fig. 94) are naturally always in the same plane, and as 
they are equal, their resultant R is a force equal to twice that 
acting on the individual throw, acting at a point midway between 
the two throws. The resultant Tvh of the centrifugal forces F 1 
acting on the two inner throws is exactly equal to the resultant 
of the centrifugal forces acting on the two outer throws, and 
acts at the same point but in the opposite direction to the latter; 
consequently it neutralizes or balances it. 

But while a crankshaft of this type is perfectly balanced as a 
whole, its individual throws are unbalanced, and since the crank¬ 
shaft is more or less flexible, the centrifugal force acting on the 
individual throws tends to produce a pressure of the crank jour¬ 
nals adjacent to these throws on their bearings. To overcome 
this, some foreign designers provide each of the short crank arms 
in a four cylinder crankshaft with a balance weight, so that the 
crankshaft is perfectly balanced at every point of its length. 

In a four cylinder two bearing crankshaft there is no inherent 
balance, since there are only two short crank arms, and both of 


THE CRANKSHAFT. 


173 



Fig. 94 .—Centrifugal Forces on Four Throw Crankshaft. 

these are on the same side of the crank. There is, however, a 
certain length of shaft between the two inner crank pins, and 
by making this of somewhat greater diameter than the crank pin, 
and by drilling out the two outer crank pins no difficulty need 
be experienced in balancing such a crankshaft without the*use 
of counterweights. 

In a six-throw crankshaft the throws are arranged in pairs, the 
two inner ones being in line with each other, as are also the 
two outer ones and the two intermediate ones, respectively. 
Each pair of throws is located at an angular distance of 120 de¬ 
grees from the other two pairs. By reference to Fig. 95, which is 
an end view of a six throw, seven bearing crankshaft, it can 
easily be seen that if all of the throws are identical, such a crank¬ 
shaft is in perfect balance. The resultant F of the centrifugal 
forces on the two throws of each pair acts at the middle of the 
length of the crankshaft. The three resultants therefore all act 
in the same plane radially outward from the centre of the crank- 



Fig. 95.— End View of Six Fig. 96— End View of Six 
Throw, Seven Bearing 1 hrow. Four Bearing. 

Crankshaft. Crankshaft 











































174 THE CRANKSHAFT. 

shaft, at angles of 120 degrees apart, and they exactly balance 
each other. 

Six throw four bearing crankshafts are made in two general 
forms. In the first, which was illustrated in Fig. 92, the two 
crank pins of each end pair are connected by a long arm whose 
centre forms a chord to an arc of 120 degrees, while the two 
pins of the central pair are directly connected together. The 
end view of this shaft, Fig. 96, shows at once that such a crank 
is not inherently balanced. The crank pins and the six short 
crank arms are in perfect balance, but the two long crank arms A 
are both on the same side of the crankshaft, and there is nothing 
to balance them on the opposite side except the short length of 
shaft between the two central crank pins, which is certainly in¬ 
sufficient. In order to secure a perfect balance, a fairly heavy 
annular flange B is sometimes provided in the forging between 



Fig. 97.—Six Throw, Four Bearing Crankshaft With 
. Dummy Journals. 


these two pins. If the portion B of the shaft between the two 
inner crank pins is equal in weight to one of the long crank 
arms the crankshaft will be perfectly balanced. 

The other form of six throw four bearing crankshaft is illus¬ 
trated in Fig. 97. In this design the four outer crank pins each 
have two short arms, the two inside arms of each double throw 
being connected by a short dummy journal. The two central 
crank pins are again directly joined to each other. It is evident 
that in this case two short crank arms between the two central 
crank pins are needed in order to give a perfect balance, but 
these two arms can, of course, be advantageously replaced by 
an annular flange between the two central crank pins. 

In recent years the tendency has been constantly toward more 
substantial crankshafts, and the author would advise that if 
any deviations are to be made from the results obtained by 
means of the equations given on the following pages, larger 
rather than smaller values be chosen. 












































THE CRANKSHAFT. 


175 


Compression 
Ratio (r). 

3 ... 

3.2 . 

3.4 . 

3.6 . 

3.8 . 

4 ... 

4.2 . 

4.4 . 

4.6 . 

4.8 . 

5 ... 


TABLE V. 


Normal 

Explosion Pressure (p) 
(Lbs. Per Sq. In. Gauge). 

. 173 

. 187 

. 205 

. 224 

. 241 

. 258 

. 276 

. 294 

. 313 

. 328 

. 345 


Single Cylinder Crankshafts—In a single cylinder crank¬ 
shaft the pressure on the crank pin may be made 1,200 pounds per 
square inch of projected bearing surface, based on the normal ex¬ 
plosion pressures given in Table V. The ratio of the crank pin 
length to diameter may be made 1.4. Consequently, if we desig¬ 
nate the length of the crank pin by / and the diameter by d, the 
projected bearing surface is 

d X 1.4 d = 1.4 d 2 . 

Since the total pressure on the bearing is 


p _ f b 2 7T } 

4 

with a unit pressure of 1,200 pounds per square inch, we have 


fib 2 7 T 

1,200 X 4 


1.4 c? 5 


from which it follows that 


.(67) 

46 

and 


/ = 


1.4 d = 


b $ 
33 


( 68 ) 


If the flywheel is placed at the driving end of the crankshaft the 
crank arm and crank journal at that end will be more heavily 
loaded than the other crank arm and crank journal, and should, 
consequently, be made larger. The crank journal diameter can be 
made equal to the crank pin diameter; the length of the journal 
at the flywheel end should be equal to twice the length of the 
crank pin bearing, and the length of the journal at the opposite 




















17G 


THE CRANKSHAFT. 


end equal to 1.25 times the length of the crank pin bearing. If 
the flywheel is located on the opposite side of the crankshaft 
from the driving end, or if it is desired, for esthetic reasons, to 
have the crankcase with its bearings symmetrical, both of the 
•crank journals may be made equal to 1.75 times the length of the 
crank pin bearing. 

The dimension of the crank arm parallel to the axis of the 
crankshaft will be called the thickness of the arm and designated 
by t, and the other dimension will be called the width of the arm 
and designated by w. If we denote the thickness of the arm at 
the flywheel and driving end by ti and that of the arm at the 
other end by U, then the following equations give suitable values 
for these dimensions: 


— 0.6 d ....(69) 

t 2 = o.^d . (70) 

w = 1.35 d . (71) 


Double Throw Crankshafts—Crankshafts with double throws 
between the crank journals are very extensively employed, this 
type being represented in all double cylinder opposed motors, 
which are widely used on light commercial vehicles, and in many 
four cylinder vertical motors. Double throw crankshafts of the 
two and four cylinder types operate under identical conditions, 
and the same equations will serve for calculating the dimensions 
of both, except as regards the dimensions of the crank journals. 
Owing to the greater distance between supports, the 
crank pins of these shafts are made of relatively larger diameter, 
hence their rubbing speed is somewhat greater and their unit 
pressure should be less. A good figure to use is 900 pounds per 
square inch (based on the normal explosion pressures given 
in 1 able \). The ratio of crank-pin length to diameter may 
be set down as 1.25. We then have 


d = 
and 


b $ 

~W 


tv* =1.25 d > 
900 x 4 


o 


(72) 


_ b -p 


30 


(73) 


For ordinary lengths of stroke the crank journals can be made 
of the same diameter as the crank pin, while for exceptionally long 
strokes they may be made % inch larger. The length of the 
journals for a double opposed motor should be 1.8 times that, 
of the crank pins calculated as above; that is, 










THE CRANKSHAFT. 


177 


_ b *s -p 


16.6 


(74> 


In this case, too, it would not be a bad idea to take, say, Y% 
inch from the length for the free end bearing and add it to the 
length of the driving end bearing, though this is not very often 
done in practice. There is a tendency to make the crank pins in 
double opposed motors rather shorter than the proportion given 
above, so as to bring the axes of the two cylinders closer together 
and thus secure a better balance. But when the pins are made 
longer tne crank shaft and its bushings will wear longer, and the 
above proportions are believed to give the most satisfactory 
results. 

In a four cylinder crankshaft with double throws between sup¬ 
ports the total crank journal length may be made slightly greater 
than the total crank pin bearing length. The three crank journals 
of the crank shaft are, however, not made of the same length. 
Tire forward and centre bearings may be made 1.25 times the 
length of the crank pin bearing, as determined by equation (73), 
and .the driving end bearing 1.75 times the length of the crank 
pin bearing. 

A double throw crankshaft has two kinds of arms, the short 
arms joining the crank pins to the crank journals and the long 
arms connecting crank pins on opposite sides of the shaft. The 
bending moment is greatest at the middle of the crank pin, 
from which point it decreases uniformly toward the points of 
support; hence it is greater in the long than in the short 
arm, and for this reason the long arm is made thicker than 
the short arm. Both sets of arms are made of the same 
width, which is generally a certain percentage greater than 
the crank pin diameter, equaling the diameter of the flange 
on the crank pin bushing. A good proportion is 


w = 1.25 d. 


( 75 ) 


The arms can then be made of suitable thickness to give the re¬ 
quired strength. In the case of the long arms the bending stresses 
are evidently of prime importance. The resistance to bending in 
the plane of the crank shaft is proportional to the width and the 
square of the thickness. The resistance to bending of the arm 
should evidently be directly proportional to the resistance to bend¬ 
ing of the crank pin, which latter is proportional to the cube of 
its diameter. We may therefore write 

m d x — zu ti 2 

where m is a constant. The average value of this constant in a 
large number of motors of which the writer has the data is 0.5. 
Consequently we may write 




178 


THE CRANKSHAFT. 


0.5 d 3 = zv ii*, 


from which it follows that 



( 76 ) 


/ 


The short crank arms are subjected to the same sort of stresses 
as the long arms, except that the bending stresses are relatively 
smaller, these arms being closer to the supports. Since the bend¬ 
ing moment varies directly with the distance from the support and 
the resistance to bending is directly proportional to the square of 
the thickness, it seems reasonable that the thickness of the short 
arm should be a certain fraction of the thickness of the long arms. 
In practice the short arms are generally three-fourths as thick as 
the long arms. Their thickness may therefore be found from 
the equation 



(77) 


Four Throw, Three Bearing Crankshaft—We will now 
calculate the dimensions and work out a design for a three bear¬ 
ing crankshaft for a four cylinder twin cast motor of 4 inch bore 
and 5 inch stroke, the explosion pressure to be 290 pounds per 
square inch. The crank pin diameter should be 

--I. 290 =r 1.8, say 1 13/16 inches, 

38 

and the crank pin length 

1.25 X i-8 = 2.25 or 2^4 inches. 

The length of the forward and intermediate journals should be 

1.25 X 2.25 = 2.81, say 2^4 inches, 
and the length of the rear bearing 

1.75 X 2.25 = 3.94, say 4 inches. 

The width of the crank arms should be 

1.25 X 1.8 = 2.25 inches; 
the thickness of the long crank arm 


= 1.14, say 1 x /$ inches; 

and the thickness of the short crank arm, 

H X 1.14 = .85, say % inch. 



The distance apart of the cylinder axes in each pair of cylin¬ 
ders is made up of the following items: 











THE CRANKSHAFT. 


17J» 

Inches 

Cylinder bore . 4 

Two cylinder wall thicknesses, inch each. J4 

Jacket space between cylinders. y 

Total . 4 

The minimum distance between the axes of the inner cylinders 
would be made up of the following items: 

Cylinder bore.Inches. 

Two wall thicknesses inch each. y 2 

Two jacket spaces y 2 inch each. 1 

Two jacket walls y inch each. iy 

Clearance space. y 2 

Total . 6 J 4 

Upon laying out the crankshaft to these centres we find that 
there is not enough room for the intermediate bearing. The dis¬ 
tance between centres required with the length of bearings cal¬ 
culated above and shown in the drawing Fig. 98, is made up of 
the following items: 

Inches. 

Crank journal length. 2^4 

Crank pin length. 2 % 

Two thicknesses of short arm with bosses, each 1 y inches. 2^4 

Total . 7}4 

This centre distance makes the distance between cylinder pairs 
iJ-2 inches. This may be more than desirable from some points 
of view, and in order to reduce it the expedient of offsetting the 
crank pin centre lengthwise from the cylinder axis is often re¬ 
sorted to. It can also be reduced somewhat by leaving off the 
bosses from the crank arms, and still more by making the crank 
from a slab, in which case the arms are of rectangular cross sec¬ 
tion, machined all over, and the crank pins and journals come 
right up to the side faces of the arms. 

Four Throw, Five Bearing Crankshaft—It might be sup¬ 
posed that as far as the crank pins and crank arms are concerned 
the same formulae should apply in this case as in that of the 
single cylinder crankshaft This does not hold true, however, 
chiefly because space in the longitudinal direction is limited, and 
since in this case considerable of it is taken up by the crank 
arms and crank journals the crank pins must be kept rather short. 
The unit bearing pressure can be made slightly greater than in 
the double throw between bearings crankshaft, using the same 
formula for the pin diameter but a length to diameter ratio of 
1.2 instead of 1.25. Therefore, the crank pin diameter 


















180 


THE CRANKSHAFT 



Fig. 99 .—Design for Two Bearing Crankshaft for Four Cylinder, 4 x 5 Motor. 







































































































































































































































THE CRANKSHAFT. 


181 


and the pin length 


d = 


7 ~ V p 


l— 1.2 d 


The dimensions of the crank arms should be the same as those 
of the small arms of the double-throw-between-supports crank¬ 
shaft. That is, the width of the arm should be 1.25 times the 
crank pin diameter, and the thickness of the arms may be found 
from the equation 



'0.28 d 3 

TV 


( 78 ) 


In the four cylinder motor the crank journals are generally made 
of the same diameter and length as the crank pins, except the 
crank journal at the flywheel end, which is made about 50 per 
cent, longer because it has to carry the weight of the flywheel in 
addition to supporting its share of the explosion pressure. In 
determining the length of bearings one of the considerations is 
the amount of space available. The cylinder or cylinder pairs are 
generally wanted as close together as possible, so as to give a neat 
and compact motor, and the centres of the adjacent crank pins 
should be the same distance apart as the axes of adjacent cylin¬ 
ders, so that the centres of the crank pins lie in the respective 
cylinder axes prolonged. In that case the connecting rod bear¬ 
ings will be symmetrical and will bear with uniform pressure 
over the whole length, which is not the case if they project more 
to one side of the cylinder axis than to the other. There is 
always more or less end thrust on a motor crankshaft, and sub¬ 
stantial thrust bearings must be provided. Where the crank arm 
is not of sufficient width to provide a liberal thrust bearing sur¬ 
face, the crank forging is often provided with segmental projec¬ 
tions on the arms near the crank journals. 

Four Throws Between Supports—In a shaft of this type 
the distance between supports is very great and the bending 
stresses are therefore very large. This calls for crank pins of 
large diameter, resulting in relatively large rubbing speeds at the 
crank pin bearing surface. On this account the unit pressure must 
be kept low. A pressure of about 700 pounds per square inch is 
giving satisfactory results. The length of the crank pin is gener¬ 
ally about equal to the diameter. Equating the quotient of the 
total pressure on the crank pin by the allowable unit pressure to 
the projected area of the crank pin, we have 






182 


THE CRANKSHAFT. 


t& TT /2 

700 X 4 

from which we find that 

d = b EjL . (79) 

30 

In this type of crankshaft the width of the crank arm is gener¬ 
ally made but little greater than the crank pin diameter. A good 
proportion is 

w-i. 2 d . (80) 

The crank arms must, of course, be much stronger than in crank¬ 
shafts with only two or one throw between supports, but since the 
most important stress is the bending stress in the plane of the 
crankshaft at the moment of explosion, and since the resistance of 
the crank arms to this stress is proportional to the square of their 
thickness and only to the first power of their width, the neces¬ 
sary strength is obtained by increasing the thickness rather than 
the width of the arm. 

The deflection of the crankshaft due to the bending in any 
part may be represented by the equation 

c Plm 
v =- 

El 

where c is a constant, P the explosion pressure, l the length of 
the member, m its centre distance from the nearest point of 
support, E the coefficient of elasticity and I the moment of in¬ 
ertia of the section. Considering the same members in crank¬ 
shafts of different sizes, m is directly proportional to the bore 
and may be written Cib. For the long crank arms, which are 

XV t ^ 

of rectangular section, the value of / is- • 


Substituting 

T2 cc l Plb 

y — -— • 


xv P E 

Hence 

P lb 


y~ 

y u> P 


But y, we found, may increase directly with the bore, hence 



from which it follows that 

PI , , 

-5= constant. 

xv P 











THE CRANKSHAFT. 


183 


This type of crankshaft is not very well standardized and the 
values of the constant C in actual crankshafts of this type vary 
widely. Assigning it q value of 2,000 will give good results. 
Hence 


t 



PI 

2000 XV 


PI 
zv t ' 6 


= 2,000 


(81) 


The thickness of the outer arm can be made two-thirds that 
of the inner arm. 


Four Throw, Two Bearing Crankshaft Design—In a block 

motor of 4 inches bore by 5 inches stroke, the distance between 
adjacent cylinder axes can be made 454 inches throughout. The 
requisite crank pin diameter in a crankshaft for such a motor is 

4 •>/ 260 _ 2 i inches, 

30 

and the crank pin length is the same. 

The width of the crank arm is 

1.2 X 2j =2.55, say 2 \ inches , 

and the thickness of the long arms is, therefore (equation 81) : 
l/ 3265 X 5 

V '—^ w- = 1.48, say 1* inches. 

r 2,000 X 2.5 ^ ’ } 1 

The thickness of the small arms is two-thirds of this, or 1 inch. 
A drawing of a four throw, two bearing crankshaft, according 
to these dimensions, is given in Fig. 99. The crank journals 
are arranged for ball bearings, which are frequently employed 
for this kind of crankshaft. If plain journals are to be used 
they can be made of a diameter equal to 0.8 times the diameter 
of the crank pin, the journal at the flywheel end being made 
twice the length of the crank pin, and that at the forward end 
1.5 times the length of the crank pin. 

It may here be pointed out that in four cylinder motors two 
bearing crankshafts are generally employed in conjunction with 
cylinders cast en bloc, three bearing crankshafts with twin cast 
cylinders, and five bearing crankshafts with individually cast 
cylinders. Three bearing crankshafts are, however, also quite 
frequently used for motors of the en bloc type, especially if the 
middle cylinders are cast some distance apart, to provide room 
for an inlet passage, for instance. 

Stresses in Members—It will be well, after having de¬ 
signed our crankshaft, to calculate the stresses in the different 
members at the time of explosion. The stresses in the crank 










184 


THE CRANKSHAFT. 


arms should be less than those in the crank pins, because a low 
stress in the crank arms means increased weight only, while a 
low stress in the crank pins means increased'friction as well as 
weight. 

To calculate the stresses we must first determine the reaction 
at the two supports. The sum of the two reactions is equal to 
the explosion pressure on the crank pin, and the individual 
reactions are inversely proportional to their distances from the 
point of application of the explosive force. The force of the 
explosion in this case is 

12.56x260 = 3,265 pounds. 

The distances from the centre of one of the inner crank pin 
bearings to the centres of the supporting bearings are 7^ inches 
and 12^4 inches respectively. (See Fig. 100.) Consequently 
the reactions at the supports are 


12I X 3,265 , , 

—--——- = 2,011 founds, 

I2f + 7s 

and 

7 i X 3 ^ 265 _ Ij2 54 -pounds , 
i2i + 7 l 

The distance from the centre of one of the outer crank pin bear¬ 
ings to the centre of the nearest supporting bearing is 3 % inches, 
and the distance to the farthest supporting bearing is 17^ inches 
Hence, when an explosion takes effect on one of the outer crank 
pins, the reactions at the supports are 

I 7 ^ ^ 3»265 _ 2,772 pounds , 

I 7t + 3i 

and 

3j ^ = 493 Pounds. 

x 7s ~r 3s 

Working now from the right hand support, we find that the 
bending moment at the centre of the right hand inner crank pin 
at the moment when an explosion takes effect on that pin is 
7 7 A x 2011 = 15,837 pounds-inches, 

and the bending moment at the centre of the right hand outside 
crank pin when an explosion takes effect on this pin is 
3 l A x 2772 = 8,663 pounds-inches. 

These bending moments, as stated, are the maxima, and the 
moments decrease uniformly from these two points to the points 
of support, where they vanish, as shown in Fig. 100, top por¬ 
tion. From this bending moment diagram it will be seen that 






THE CRANKSHAFT. 


185 


the outer end of the crankshaft up to the point E sustains the 
greatest bending effect when there is an explosion on the outer 
crank pin, while the inner portion from the point E to the 
centre is subjected to the greatest bending effect when there is 
an explosion on the inner crank pin. The four points for which 
we have to determine the greatest bending stresses are the centre 
of the short crank arm (A), the centre of the outer crank pin 
( B ), the farthest point along the centre line of the long crank 
arm from the right hand support where it still retains its regular 
cross section (C), and the centre of the inner crank pin ( D ) 
The bending moments at these four points are 



Fig. ioo. — Bending Moments in Four Throw, Two Bearing 

Crankshaft. 

At A —1^x2772= 3,638 pounds-inches. 

At B — 2772= 8,663 pounds-inches. 

At C —5Ms x 2011 = 10,306 pounds-inches. 

At D —7% x 2011 = 15,837 pounds-inches. 

In order to find the stresses under which the material at these 
points works at the moment of explosion we equate the resisting 
moments to the bending moments found above as follows: 

At^- 2 - 5XlX ^ = 3,638 
6 

8,731 -pounds Per square inch. 

MB- 3,14 X 2.125 s X S = 8 663 


32 













































186 


THE CRANKSHAFT. 


S = 9,200 poufids fer square inch. 

At C— 2-5 X I, 5 *X ^—,0,306 
6 

6" = 11,000 founds fer square inch. 

At D - 3 ^4 X 2.125 8 X S = l5 g 37 
32 

5=16,820 pounds per square inch. 

It will thus be seen that the stress in the inner crank pins is 
about 50 per cent, greater than the stress in the long crank arms. 
The idea suggests itself that since the stress in the outer crank 
pin is little more than half as great as in the inner crank pin, 
the outer crank pin may be made smaller in diameter. This 
is not done in practice because it is not desirable to have two 
•different sizes of connecting rods. In the present design, if the 
outer crank pin were to work under the same stress as the 
inner one its diameter would be about 1 inches, which would 
mean a reduction of about one-sixth in the friction loss on these 
two pins—hardly enough to justify the use of two different sizes 
•of connecting rod, besides the slightly reduced stiffness of the 
crank. 

Types of Six Throw Crankshafts—Six throw crankshafts, 
which in their final shape have their throws spaced 120 degrees 
apart, are forged with all their throws in a single plane. After 
the forging has been completed, and while it is at a red heat, 
the four end throws are bent through such angles as to bring 
adjacent crank pins 120 degrees apart. In order to facilitate this 
operation a small dummy journal is sometimes provided between 
the first and second and between the fifth and sixth throws. 
It has been found that crankshafts for six cylinder motors must 
be made very rigid, otherwise there is trouble from torsional 
vibration. In the formulae for crank pin diameter 

i= h JjL 

C 

the value of c may be taken at 34 for seven bearing, 32 for four 
bearing and 30 for three bearing shafts, and the ratios of crank 
pin length to diameter at 1 for seven and three bearing shafts 
and 1.1 for four bearing shafts. The journal bearings, except 
the rear one, will be somewhat shorter than the crank pin in a 
seven bearing crankshaft, slightly longer than the crank pin 
in a four bearing crankshaft and about 1.3 times as long in a 
three bearing crankshaft The arms can be proportioned with 
the aid of the rules already laid down for other types of crank 
.shafts. 





THE CRANKSHAFT. 


187 





A design for a crankshaft 
for a six cylinder 4x5 inch 
motor is given in Fig 101. The 
two long crank arms are there 
shown as running straight 
across from pin to pin. Often 
they are made to curve in to¬ 
ward the axis of the shaft, for 
the sake of balance. 

Front and Rear Ends— 
The front end of the crank¬ 
shaft must be extended to car¬ 
ry the driving pinion of the 
camshaft gear and the pin or 
ratchet toothed sleeve with 
which the starting crank en¬ 
gages. There may also be an 
oil guard formed on the shaft 
inside the seat for the cam 
gear pinion. The length of the 
cam gear seat may be made 
*4 inch per inch bore of the 
motor and the length for the 
starting ratchet seat the same. 
The starting ratchet seat is 
made smaller in diameter than 
the pinion seat, and the latter 
smaller than the crank journal. 

At the rear end of the crank¬ 
shaft an oil guard is usually 
provided, especially if a cone 
clutch is to be employed. This 
consists of a narrow flange 
turned on the crankshaft just 
beyond the rear crank journal, 
with a sharp outer edge, from 
which the oil is flung off when 
the shaft turns rapidly. The 
oil is caught in a recess turned 
in the outer end of the bear¬ 
ing hub, from which it drains 
back to the crank case. 





















































































188 


THE CRANKSHAFT. 


Beyond the oil guard there is generally provided an integral 
flange to which the flywheel is bolted. This flange is made of 
an outside diameter equal to the stroke, of a width equal to 
% inch per inch of bore, and is drilled with six holes for the 
reception of bolts. The diameter of the bolts may be made 
equal to the width of the flange. Where a very secure job is 
desired the holes in the flange and in the web of the flywheel 
are counterbored and have hardened dowel bushings inserted, 
which take up the driving strain. 

Beyond the flywheel flange the crankshaft is generally ex¬ 
tended a short distance to form a stud for the clutch to ride 
upon. 

Crankshaft Blanks—Crankshafts are made either from large 
slabs of steel of prismatic form slightly larger than the over¬ 
all dimensions of the finished shaft, or from drop forgings, which 
are very close to the sizes of the finished shaft. When the shaft 
is made from a slab it is practically impossible to provide an 



Fig. 102.—Slab of Steel Drilled to Form Cranks. 


integral flywheel flange, and the flywheel is then keyed on, a long 
tapered key seat and a thread for a nut to draw the flywheel 
onto its seat being provided for the purpose. The cranks are 
first formed by drilling and sawing, as indicated in Fig. 102, and 
are then rough turned, in which state they are generally deliv¬ 
ered to the motor manufacturer. Crankshafts of this type are 
eventually machined all over, being planed or milled on the flat 
sides and turned on all other surfaces. 

When the cranks are to be produced by the drop forging 
process a draft of about 8 degrees must be allowed on the 
sides of the arms, so the forgings can be readily withdrawn from 
the dies. The sections should be laid out so that they are equiva¬ 
lent to rectangular sections of the calculated dimensions. When 
the forgings are completed they are subjected to the heat treat¬ 
ment already referred to, and are then ready for machining. 

Machining—Whether the crankshafts are made from slabs 
or drop forgings, they are generally first turned in the lathe, 










THE CRANKSHAFT. 


189 


rnd the crank pin and journal bearings are then finished in a 
^rinding machine. It was formerly customary to finish them in 
the lathe, but since a four or six cylinder crankshaft supported 
only at its ends is not very stiff, it tends to spring away from 
the cutting tool, thus becoming inaccurate. In grinding, on the 
other hand, the pressure of the wheel against the work in taking 
the finishing cut is so light that there is no springing away, and 
an almost absolutely round journal is produced. 

For turning the crank pins the crank shaft must be’ secured 
in the lathe in such a manner that the crank pin axis coincides 
with the lathe centre axis. After the pins in one angular position 
have been finished the crankshaft must be turned around so as 
to bring other crank pins into the axis of the lathe centres, to be 
machined, and the angular movement given the crankshaft must 
be very accurately gauged in order that the crank shaft may 
be absolutely true and the crank and piston motion properly bal¬ 
anced. Moreover, since the centre of gravity of the crankshaft, 
•when the latter is so supported that one of its pins is in line with 
the lathe centres, is at distance from the centre line of rotation, 
the lathe face plates or crank clamping fixtures must be pro¬ 
vided with counterweights to balance the weight of the crank¬ 
shaft, so the lathe may be run at high speed without undue 
vibration. 


A simple clamping fixture for a four cylinder crankshaft is 
shown in Fig. 103. The shaft is clamped by its outer journals 
in the middle of the clamping 
fixture. The latter is provided 
with two centre holes for the 
lathe centre, one in line with 
each pair of crank pins. With 
this fixture suitable weights for 
balancing the weight of the 
crankshaft must be clamped to 
the face plate, and spacers must 
be placed between crank arms 
to prevent their being forced 
together under the pressure of 
the lathe centres. 

A crankshaft turning lathe is 
illustrated in Fig. 104. The 
crankshaft is clamped in clamp¬ 
ing fixtures by means of swing¬ 
ing caps, and is held in the proper angular position by a 
sliding Vee carried on the fixture secured to the head stock. 



Fig. 103.—Simple Crank 
Turning Fixture. 















190 


THE CRANKSHAFT 



Fig. 104.—Crankshaft Turning Lathe Made by the American Tool Works, Cincinnati, Ohio. 








THE CRANKSHAFT. 


191 


\\ hen one set of pins has been machined, a portion of the fix¬ 
ture can be swung through a half circle in order to bring the 
other set of crank pins in line with the lathe centre. The tool 
at the front of the lathe carriage is used for turning the cylin¬ 
drical or parallel surface of the crank pins, while the two tools 
at the rear end of the carriage are used for turning the fillets, 
which should always be provided at the junction of crank pins 
and journals with crank arms. 



Fig. 105.— Crankshaft Grinding in a Machine Made by thi 
Landis Tool Co., Waynesboro, Pa. 

In Fig. 105 is shown a portion of a crank shaft grinding 
machine with the crankshaft in place. The latter is held at both 
ends by quickly operated clamping holders, and is driven from 
both ends, the head and tail stock of the machine being geared 
together. There is an eccentric adjustment of the work carrying 
heads on the face plates, by means of which the different throws 
can be brought into line with the spindle centre, and in this 
particular machine the setting is indicated by either English or 
metric scales. For multiple throw shafts the pins are ground in 
their exact position relative to each other, being located by in¬ 
dexing fixtures on the heads which have twelve 30 degree divi¬ 
sions, which give all the regular throw angles. 

Hollow Crankshafts—In a number of motors the crank 
bearings are lubricated by feeding oil through the centre of the 






192 


THE CRANKSHAFT. 


crankshaft, and the latter must then be drilled through the crank 
journals, arms and pins, and the ends of the drill holes plugged. 
A typical shaft of this kind is the Marmon, illustrated in Fig. 
106. The oil enters the channel in the crank shaft through 
radial holes drilled half way through the outer journals, which 
register once during every revolution with holes in the crank 
bushings and bearing hubs, to which oil is being fed by a pump. 
The oil travels through the horizontal passage in the crank jour¬ 
nals under the head produced by the oil pump, and is carried out 
into the crank pins by centrifugal force. The crank pins are also 
drilled with radial holes at their middle, through which the oil 
flows to the bearing surface. Some makers take advantage of 
the centrifugal force to feed oil from the crank pin to the piston 
pin bearing by connecting the lower connecting rod bearing with 
the upper by means of a tube running either through or along¬ 
side of the connecting rod. 

Ball Bearing Crankshafts—Although the great majority of 
crankshafts are provided with plain or parallel journal bear¬ 
ings, the so-called radial type of ball bearing has recently met 
with increased application to crankshafts, especially to crank¬ 
shafts of the four throw, two bearing type. Perhaps the prin¬ 
cipal advantage of the radial ball bearing for this purpose is 
that it permits of reducing the over-all length of the motor, and 
since this is one of the main objects aimed at in en bloc cylinder 
construction, and the use of four throw, two bearing crankshafts, 
the ball bearing falls right in line with this class of design. One 
further advantage of ball bearings on crank shafts is that they 
reduce the friction loss. Of course, ball bearings are not at the 
same advantage in a place where the load varies between wide 
limits as where it is fairly uniform, and bearings of relatively 
large size must be chosen in order to secure satisfactory results. 

In applying ball bearings to crankshafts which have journals 
between throws, some difficulty is encountered in getting the in¬ 
termediate bearings into place. In order to avoid this difficulty 
some designers use ball bearings at the ends only, and plain 
bearings in the middle. Where ball bearings are to be applied 
throughout, the inner bearings must be chosen of such a bore 
that they can be stripped over the crank arms. To facilitate this 
process the ends of the crank arms are rounded off so that the 
bore of the bearings may be made smaller. 

Practically all the radial ball bearings used in this country are 
imported, and are made to metric measurements. These bearings 
come with bores of the inner races in 5 millimeter gradations. 
For a four throw, three bearing crankshaft a ball bearing of a 


THE CRANKSHAFT 


193 



(A 

> 

< 

£ 

j 

C 


< 

£ 

3 £ 

W 

£ 


H 

f-T 

fc 

< 


cfl 

£ 

<; 

u 

£ 


PS 


'g 

►— 


ci 











































































































































































194 


THE CRANKSHAFT. 


bore equal to the crank journal diameter will be big enough for 
the front end; a bearing that can be passed over the crank arms 
will be big enough for the central journal, and a bearing of the 
same size as this will do for the rear or flywheel end. This 
means, of course, that the intermediate and rear bearings have 
to be fitted to enlargements of the crank journals or to collars 
surrounding the journals, which latter, in the case of the inter¬ 
mediate bearing at least, have to be made in halves in order to 
get them into place. 

These radial ball bearings are made in three different weights; 
that is, for a certain bore of inner race the bearings are made 
with three different sizes of balls and outer diameters. For the 
flywheel end, if the flywheel is secured by a key so the bearing 
need not be stripped over the crankshaft, it is advantageous to 
use the heavy type of ball bearing, in which case the crank 
journal need not be enlarged, or at least not much, at the bear¬ 
ing seat. 

As already stated, ball bearings are used most extensively on 
four throw, two bearing crankshafts. One advantage of their 
use on such shafts is that they permit considerable flexure of 
the shaft without imposing extra strains upon it. The sizes of 
bearings to select should be discussed with the maker of the 
bearings. The flywheel end bearing should, of course, be some¬ 
what larger than the front end bearing. In one design the fly¬ 
wheel end bearing is mounted on the hub of the flywheel, where¬ 
by the motor is rendered more compact. 

Fig. 107 is an illustration of the crankshaft of the Stearns 
30-60 horse power motor, which is mounted in three radial ball 
bearings. This is a four cylinder motor of 5inches bore and 
5 % inches stroke. The front bearing is of the heavy type of 2.56 
inches bore, while the intermediate and rear bearings are of the 
medium weight type, of 3.15 inches bore. The rated load capacity 
of each of these bearings is approximately 5,000 pounds. 

Testing the Running Balance of Crankshafts—Fig. 108 
shows a machine for testing the running balance of crankshafts 
and other parts, made by the Norton Grinding Co., of Worcester, 
Mass. The crankshaft is rotated by means of an electric motor 
secured to the frame of the machine, while carried on yielding 
supports, and if it is not in running balance it will cause the sup¬ 
ports to vibrate in the horizontal plane, the vibrations being indi¬ 
cated by means of long indicating hands. 

Referring to the illustration, the machine consists of a sta¬ 
tionary support with three upright, lengthwise adjustable 


THE CRANKSHAFT. 


195 


columns A which carry vertical steel rods B, fitted at their 
top end with double pairs of rollers C each, on which the 
crankshaft may rest. The steel rods B are freely supported 
at their lower ends in brackets D with cup-shaped recesses, 
and are flexibly held near the top ends by means of rubber 
discs carried by the upright columns. This allows the top 
ends of the rods to move to and fro under the action of the 
centrifugal force on any unbalanced part of the crankshaft. 

The crankshaft is driven by means of a belt from a hori¬ 
zontal shaft which receives its motion from the electric 
motor through an intermediate vertical shaft with universal 



Fig. 108.—Norton Crankshaft Balancing Machine. 

joints, a friction drive mechanism and a pair of helical gears. 
An‘'adjustable scriber E is fixed to the top of each vertical 
column, and the top of each carrying rod B is connected to 
a long indicator hand F extending downward at the side of 
the column, so that the vibrations of the rod are multiplied 
and plainly indicated. In the use of the machine the crank¬ 
shaft is rotated at a certain speed and the different scribers 
are advanced until they mark the shaft; then the scribers 
are slightly withdrawn, the direction of rotation is reversed 






196 


THE CRANKSHAFT. 


and the scribers are again advanced until they mark the 
shaft. The part of the shaft midway between the two marks, 
circumferentially, is too heavy and should be lightened. A 
color roller is furnished with the machine, by means of which 
color is applied to the revolving journal to make the scriber 
marks plainer. 

After the crankshaft has been properly balanced at mod¬ 
erate speed it can be tested at service speed. It is claimed 
that the indicating means is so sensitive that a proper balance 
can be obtained by running at low speed only, in case the 
crankshaft is not sufficiently stiff to permit of driving it at 
high speed without causing it to flex appreciably. 





CHAPTER X. 


THE CONNECTING ROD. 

The connecting rod is the intermediate member between the 
crankshaft and the piston. It consists of the rod proper, the 
head or big end, which forms a bearing for the crank pin; and 
the small end, which forms a bearing for the piston pin. 

Materials—Connecting rods are almost invariably drop forged 
from either low carbon (about 0.30%) or nickel steel. The for¬ 
mer may be assumed to have a tensile strength of 70,000 pounds 
per square inch and the latter of 90,000 pounds per square 
inch. For heavier work they are occasionally cast of steel or 
manganese bronze. 

Length of Rod—In determining the dimensions of the rod 
proper we have to consider first the necessary length, and, second, 
the necessary cross section. In speaking of the length of the 
connecting rod the distance between the centre of the piston pin 
bearing and the centre of the crank pin bearing is referred to. 
This length is generally expressed in terms of the stroke. In our 
discussion of the crank moment and of side thrust on the cylinder 
wall we assumed different values for the ratio between the length 
of the connecting rod and the length of stroke ranging between 
1.75 and 2.5. 

For a motor with offset cylinders and intended for a purpose 
where lightness is paramount, the smaller ratio can be chosen. 
In American pleasure car motors of the symmetrical type the 
ratio lies generally between 2 and 2.25. European designers not 
infrequently provide a connecting rod equal to 2.5 times the 
length of the stroke and this ratio can be recommended for 
commercial vehicle motors in which long life of the motor is 
more important than low weight. 

A consideration that influences the length of the rod is that 
it must clear the bottom end of the cylinder bore when in its 
position of greatest angularity. The cylinder, crank and con¬ 
necting rod are laid off on the drawing board, as shown in Fig. 


197 



198 


THE CONNECTING ROD. 


109, with the connecting rod at right angles to the crank arm. If 
a certain length of connecting rod does not clear properly, a longer 
rod may be chosen and the layout worked over again. Some¬ 
times when the shortest possible connecting rod is wanted the 
bottom end of the cylinder is formed with slots into which the 
connecting rod swings at the moments of its greatest angularity. 

Calculation of Cross Section —The connecting rod works 
chiefly under compression, and must be treated as a column. 



Fig. 109. 



B 

Fig. no. 


The most accepted formula for the strength of columns seems 
to be Rankine’s, which is as follows: 


P 

A 


5c 

1 + Q r l 


(82) 


where P is the total pressure on the column; A, the cross sec¬ 
tional area at the section of maximum stress; Sc, the working 
stress; / J the square of the length of the column; r 2 the square 
of the least radius of gyration and q a coefficient depending upon 
the condition of the ends of the column (whether fixed or free) 
and upon the material. The square of the least radius of gyra¬ 
tion is found by dividing the moment of inertia of the 
section by its area. The left hand member of the equation evi¬ 
dently represents the compression per unit of area of the cross 





































THE CONNECTING ROD. 


199 


• / 2 
section of the column. The term q —- in the denominator of 

r 1 

the right hand member of the equation takes account of the 
additional stress due to the bending of a long column. When 
the length of the column is small in comparison with its cross 
section this factor is negligible, and the maximum stress is 

simply — • 

A 

In the discussion of columns with respect to their strength, 
a distinction is made between columns fixed at their ends and 
those free at their ends, the latter being not nearly as strong 
as the former, if the length is large in comparison with the cross 
section. Now, the connecting rod presents a case intermediate 
between a column with fixed ends and a column with free ends, 
since it is free at its ends in the plane of motion, and fixed at 
its ends perpendicular to the plane of motion. The result is that 
if bending occurs in the plane of motion it will be as shown 
on an exaggerated scale at A in Fig. no, the entire length of 
the rod tending to form a simple curve concave to the normal 
centre line of the rod. If bending occurs at right angles to the 
plane of motion, the rod will assume a shape similar to that 
shown at B, Fig. no. The end portions of the rod will form 
curves convex to the normal centre line, and the centre portion 
a curve concave to the normal centre line. At the points where 
the convex and concave portions join there is a neutral section 
at which there is no bending stress. The distance between 
these two inflection points is equal to one-half the total length 
of rod, and since the stress necessary to cause a certain bending 
strain varies as the square of the length of the curve, it follows 
that the bending moment due a certain pressure on the rod is 
four times as great in the plane of motion as normal to the 
plane of motion. 

In a connecting rod we have, however, an additional moment 
tending to cause the rod to bend in the plane of motion which is 
not present in a stationary column—namely, the moment due to 
the transverse acceleration of the mass of the rod, or the whip¬ 
ping effect, as it is generally called. But at the instant when 
the greatest pressure acts on the connecting rod. the lower end 
of the rod moves transversely at a uniform speed, and the trans¬ 
verse accelerating force is zero. This effect may, therefore, be 
neglected. 

The formula quoted above, viz.: 




200 


THE CONNECTING ROD. 


may be written 


p i n p _ c 

— -f q —, — = 5 C . 

A r* A 


( 83 > 


which shows that the actual maximum stress on the fibres of 
the rod is made up of two items, the first, * being that due to 

the compression of the rod and the second, q a — , being that due 


to the bending moments. 

From a mass of practical data at hand the writer has calcu¬ 
lated the proper values for q and for 5 c in equation (83). The 
value of q is 0.000526, considering the square of the least radius 


y 





nr 

T 


,— W4t- 

1 

-t- 

1 

* 

K> »0 



! 

t 



—3.8t- 



Fig. 112. 


of gyration around the axis x-x, Fig. in, and that of 5 c, I 3 >°°° 
in the case of low carbon drop forge steel and 16,700 in the case 
of nickel steel. 

Hence we may write 

— -f-0.000526 - ^-—= 13,000. .(84) 

A A r* 

for carbon steel and 
P PI a 

— -j- 0.000526-- = 16,700.(85) 

A A r 2 

The section of the connecting rod is generally of I shape and 
the flanges must be given the usual draft. The writer has found 
it convenient, in calculating the dimensions of the section, to base 
the calculations on an equivalent section with web and flanges ol 
uniform thickness, that is, without draft. An analysis of the 
sections in actual use shows that the average proportions are as 
follows: Calling the thickness of the web and the flange t, the 
width of the section is 3.8 and the height is 5.7. The height is, 

































THE CONNECTING ROD. 201 


therefore, one and one-half times the width. The moment oh 
inertia around the axis x-x (Fig. 112) is 

A= 3 ^ <( 5 - 7 <) , - 2 - 8 <( 3 - 7 <y ,= 46 . 8 3 < . f 

12 

and the moment of inertia around the axis is y — y — 

j _ 2/( 3 - 8 /) 3 + 3 * 7 '(' 3 ) 

Jy - 9*45 1 • 

12 

The area of the section is 

(3.8/ X 5.7 t ) — (3.7/ X 2.8 /) = 11.3 t 2 . 

Since the square of the least radius of gyration around any axis 
is equal to the moment of inertia around that axis divided by the 
area of the section, we have for the square of the least radius of 
gyration around x — x 


46j_8_3 A 

11.3^ 


4.14/ 2 


and the square of the least radius of gyration around y — y 


11.3 t 2 

Having thus found expressions for A and r x 2 in terms of t, we 
may substitute these values in equations (84) and (85) and 
solve for t. Equation (84) becomes 


P PI 2 

-- +0.000526- - - = 13,000 

11.3^ n-3* 4* J 4 * 

4.14 P t 2 + 0.000526 PI 2 = 608,140 i * 
t K — o.0000068 Pt 2 = o . 000000000866 PI 2 

l 2 — o. 0000034 P — ^ ( 0 • 0000034 P ) 2 + o. 000000000S66 P l 2 



This equation, which gives the necessary thickness of web 
and flanges for a section of a carbon drop forge steel connecting 
rod of length l and subjected to an explosion pressure P, is 
rather inconvenient to use on account of the big fractions in¬ 
volved, and for this reason Chart III has been drawn from 
which the proper value of t for any values of P and / can readily 
be determined. A formula similar to (86) has been worked out 
for nickel steel and this is represented by a separate line in the 
chart, so that the proper thickness of section for either carbon 
or nickel steel can be determined. After the thickness has once. 














0.25 0.24 0.23 0.22 0.21 0.20 0.19 0.18 0.17 0.16 0.15 0.14 0.13 0.12 In 


202 


THE CONNECTING ROD. 


been found the width of the section can be found by multiplying 
the thickness by 3.8 and the height by multiplying the thickness 

by 5-7* 

The section thus calculated is to be used at the middle of the 
length of the rod where at the moment of explosion the bending 
stress is the greatest. It is the usual custom to taper the rod, 
letting the height of the section decrease more or less rapidly 
from the head down to the small end. There is no apparent 



















































































































THE CONNECTING ROD. 


203 


reason why this should be done, since the maximum stress occurs 
midway between the ends. The theoretically correct way would 
be to make the section largest at the middle and taper it slightly 
toward the ends, but the gain would be so slight that it would 
not pay for the more complicated die. The writer considers a 
straight rod of uniform section from end to end the most prac¬ 
tical form. But, as stated, the ijsual practice is to taper the rod 
from the head end down to the smaller or top end. 

The form of cross section depicted in Fig. hi and Chart III 
corresponds in proportions to the mean of modern practice. If 
it is desired to use an I section differently proportioned, it is, 
of course, quite permissible to do so, but the bending moment 
around either principal axis should be at least as great as 
that of the section found from Chart III for the particular ex¬ 
plosion pressure P and length of rod l. 

Whipping Effect—It is of interest to investigate the “whip¬ 
ping effect” in a connecting rod, which has evidently a con¬ 
siderable influence on the stress in the rod when the piston is 
near mid-stroke. In Fig. 113, let A represent the momentary 
position of the crank pin axis and B C the linear speed of the 
crank pin. If we call the number of revolutions per minute N, 
then the linear speed is 


- feet fer second. 

30 X 12 

The linear speed can be resolved into two components, the verti¬ 
cal component O B and the transverse component 0 C. It is 
the latter component that interests us here. But 

O C = B C cos B C O = B C cos BOA, 
because O A is perpendicular to B C and O B perpendicular to 
O C, hence angles B C O and BOA are identical. Substituting 
the value of O C and denoting angle B O A by d as in formei 
discussions, we have for the transverse component of the crank 
pin velocity 


IT r A T a 

v — -cos ft . 

360 

To find the transverse acceleration we differentiate this ex¬ 
pression with respect to time, which gives 

dv nr If ,,- add 
-= —-sin v — • 

d t 360 dt 

But 


d 0 7 r A r 

_ _ * 

d t 30 

hence 


acceleration = 


7r 2 N 2 r 


sin d feet per second per second. 


10,800 








204 


THE CONNECTING ROD. 


Now, it is only the body portion of the rod that causes a 


whipping effect, since the 
two hubs are symmetri¬ 
cal around their journals 
and therefore balance 
around the supports, as 
it were. But we will not 
be far wrong in assum¬ 
ing that the rod extends 
from centre to centre of 
bearings. This will give 
a result slightly too 
large, so any calculation 


B 



O 


Fig. 113. 


of dimensions based upon it will be on the safe side. 

The above expression for the acceleration attains its maximum 
values when S is 90 degrees and 270 degrees, when sin 
0 is equal to 1 and — 1, respectively. The maximum values of 
the transverse acceleration are therefore 



10,800 

This is the maximum transverse acceleration of the lowest 
point of the connecting rod. The uppermost point of the rod has 
no transverse motion, and consequently no transverse acceleration. 
The maximum transverse acceleration of any intermediate point 
at a distance x from the upper end is 

x 

— am , 

where 1 is the centre to centre length of the rod and a m the 
maximum transverse acceleration of the lower end, Let s repre¬ 
sent the cross sectional area of the rod, which we will consider 
to be constant from end to end. Then the weight of a small 
section d x of the rod is 

W =0.26 s d x pounds, 
and the acceleration force on it is 

IV a 


^ jv 7r 2 N 2 r 
l 10,800 


0.26 s d x X — 


32.16 

N 2 r s x d x 
135 . 320 / 


Integrating between the limits x-o and x = l, we have for the 
total force of acceleration 


N * rsl 
270,640 










THE CONNECTING ROD. 


205 


In applying this for¬ 
mula to a practical case 
it is well to give N the 
value corresponding to 
the extreme racing speed. 

This is about twice the 
normal speed and for a 
4x5 inch motor may be 
assumed to be 2,400 r. 
p. m. From Chart III 
and the equation for the 
area of the cross section we find that the proper area 
for a rod 10 inches long and supporting an explosive force of 
3,500 pounds is 0.35 square inch. The value of r is 2.5 inches 
and that of /, 10 inches. Hence we have for the force of trans¬ 
verse acceleration on the rod 

2,400X2,400X2.5X0.35X10 0 , , 

—------—-= 187 founds. 

270,640 

When considered with respect to this force the rod obviously 
constitutes a simple beam subjected to a non-uniformly distributed 
load. The sum of the reactions at the supports is equal to the 
total pressure, viz., 187 pounds. Besides, the two reactions will 
bear to each other the same relation as if the rod was subjected 
to a concentrated load at the centre of gravity of the distributed 
load. In Fig. 114 the distributed load is represented by down¬ 
wardly pointed arrows bearing against the base line which rep¬ 
resents the length of the connecting rod. This distributed load 
can be balanced by a concentrated load of 187 pounds ap¬ 
plied at a point one-third the length of the connecting rod from 
the lower end, since the distributed load is represented by a 
triangle, and the centre of gravity of a triangle is at one-third 
its height from the base line. But a reaction of 187 pounds at 
one-third the connecting rod length from the lower end is easily 
resolved into reactions of 125 pounds at the lower end, and 62 
pounds at the upper end of the rod, as shown in the figure. 

We must now find the maximum bending moment on the rod 
due to this distributed force of acceleration. Referring to Fig. 
115, assume any point at a distance x from the upper support. 
The bending moment at this point is equal to the moment of the 
reaction at the left hand support around the point, minus the 
moment of that part of the distributed force to the left of 
this point, around the point. The moment of the reaction at the 
left hand support is evidently equal to 62X. In order to find 
the moment of that portion of the distributed force to the left 





















206 


THE CONNECTING ROD. 


of the point, we take a section dy to the left of the point x. 
Since the total length of the rod is io inches the average pressure 
on it is 187 f 10= 187 pounds per inch in length, and the maxi¬ 
mum is twice this, or 37.4 pounds per inch. The pressure at a 

distance y from the left hand support is y pounds per inch in 

10 

length, and the pressure on a short length dy at that distance 
from the left support is 

37-4 X — c ? y = 3 . 7 iydy . 

10 



The moment of this force around a point at a distance x from 
the left hand support is 

3-74 y dy (x — y) = 

3.74 (xydy — y 2 dy) 

Integrating between the limits y = o and y — x we get the expres¬ 
sion 

3 ‘ 74 ^3 — 3 _iZ 4 x 3 = 0.62 x 3 . 

2 3 

Hence the bending moment around a point at a distance x from 
the left hand support is 

M = 62 x — 0.62 x 3 . 

In order to find the point where the bending moment is a 
maximum we place the first differential coefficient of the ex¬ 
pression equal to zero— 

62 — 3 x 0.62 x 2 = o, 

from which we find 

x = 5-77 inches . 

Inserting this value of x in the equation for the bending moment 
we have 



















THE CONNECTING ROD. 


207 


M - (62 x 5.77) — (0.62 x 5.77 s ) = 238.6 pounds-inches. 

From Chart III we find the thickness f of a section proper for 
a connecting rod 10 inches long and subjected to an explosion 
pressure of 3-500 pounds to be 0.185 inch, and since the moment 
of inertia around the principal axis perpendicular to the plane 
of motion is 

7 X = 46.8 t\ 

its value in this case is 

46.8 xo.i85 4 = 0.052. 

The distance c of the outermost fibre from the neutral section is 


5.7X0. 185 
2 


= 0.527 inch . 


Hence, since the stress 

S-«i, 

I 

S — g_ 3S ; 6 X °- 5 2 7 _ 2,40 <0 pounds 
0.052 

per square inch, approximately. 

This stress must, of course, be added to that due to the com¬ 
pression on the rod. When the crank during the power stroke 
is in the quarter position the piston is 2.8 inches from the top 
end of the stroke (see Fig. 13), and the gas pressure in the 
cylinder then is 65 pounds per square inch (see Fig. 76). The 
total pressure on the piston then is 

12.56 x 65 = 816 pounds. 

Owing to the angularity of the connecting rod the pressure along 
the rod is greater in the proportion 


viz 


•) 



4 


1-031, 


1.031 x 816 = 841 pounds. 

The cross sectional area of the rod we found to be equal to 
0.35 square inch, hence the unit stress due to the compression is 


—i— — 2,400 pounds per square inch approximately. 

0.35 

The combined stress, therefore, is 

2,400 + 2,400 = 4,800 pounds per square inch 
which is very low compared with the 13,000 pounds per square 
inch to which the rod is subjected at the moment of explosion. 
This calculation shows that with steel I section connecting rods 
there is absolutely no danger from the whipping effect, even at 
the very highest speeds. 







•208 


THE CONNECTING ROD. 


-0.5 b 



Piston Pin Bearing—The upper end of the connecting rod 
generally is a plain hub bored out to receive the bearing bushing. 
The outer end of the hub can be made equal in diameter to the 
piston bosses, or 1.5 times the diameter of the piston pin, the 
bushing being made of a thickness equal to one-eighth the piston 
pin diameter and the wall of the hub at the outer end the same. 
The hub is generally made one-eighth of an inch shorter than 

the distance between the 
piston bosses, so there 
will be no difficulty in 
assembling even if the 
cylinders should not be 
fitted to the crank case 
with absolute accuracy. 
If splash lubrication is 
used the bearing hub is 
provided with a sort of 
oil pocket to catch the 
oil which drips from the 
piston head, one or 
two holes being drilled 
through the bottom of 
the pocket and through 
the bushing, so that the oil can flow to the bearing surface. 
In order that the bushing may not turn in the hub, and thus 
the oil holes through the latter be closed up, it must be se¬ 
cured by a screw or pin. When the piston pins are lubricated 
by means of an oil tube leading up from the connecting rod head 
or through their bore, this oil pocket can be dispensed with. A 
typical design of connecting rod small end is shown in Fig. 116. 

The bushing is generally made of phosphor bronze, though 
sometimes it is made of steel and hardened, in which case the 
wearing surfaces are hardened steel on hardened steel. It is 
forced into the bearing hub under pressure. Sometimes the 
piston pin bearing hub is provided with a lug and split, and a 
stud fitted with a nut and check nut is inserted, by means of 
which the bearing can be adjusted for wear (see Fig. 117). 

Connecting Rod Head—The connecting rod head is made in 
halves, with a detachable cap and half bushings, so that the bear¬ 
ing can be adjusted for wear. In engines which have large in¬ 
spection holes in the crank case the cap is sometimes hinged to 
the rod. A typical rod of this type, which is used more particu¬ 
larly on single and double cylinder motors, is illustrated in Fig. 
jt8. The advantage of the hinged cap is that it can De more 































THE CONNECTING ROD. 


209 


readily adjusted than a cap held on by two or four studs, it being 
only necessary to adjust a single screw. On the other hand, it 
seems that a bearing cannot be adjusted as accurately if pro¬ 
vided with a hinged cap as if provided with a stud-retained cap. 
The hinged bearing is divided in a plane making an angle of 25 
to 45 degrees with the cross section of the rod, to make the 
clamp bolt more accessible through the handhole in the side of 
the crank case. 

When the connecting rods are drop forged the heads are forged 




integral, and are later split by means of a saw so as to detach the 
caps. This facilitates the operation of boring out the heads. The 
heads are given the usual drop forging draft on the outside and 
are provided with bosses for the studs or bolts. Shims of thin 
sheet copper are placed between the connecting rod head and cap 
for purposes of adjustment. 

The length and diameter of the collecting rod head bore are 
fixed by the dimensions of the crank pin and need not be dis¬ 
cussed further. The thickness of the bushing can be made d/ 10, 
where d is the crank pin diameter, and the thickness of the hub 
wall at the outer ends the same, choosing the nearest size in 
thirty-seconds of an inch. 

A number of designs of connecting rod heads are shown in 
Fig. 119. In the first of these, A, the cap is secured by two 
bolts with round heads, part of which is cut away so as to pre¬ 
vent the bolts from turning when the nuts are drawn up. The 
latter are secured in place by split pins. Design B is similar, ex¬ 
cept that cap screws are used instead of bolts and nuts. These 
cap screws must be locked by some means, as by a wire passed 
through the heads of both screws with its ends turned over, or 
by a brass strip under the heads, provided with lips at its ends 





























210 


THE CONNECTING ROD. 


which are turned up against the faces of the screw heads, thus 
holding it from rotating. The feature of the design shown at C 
is that the lugs for the cap bolts are lightened by removing some 
of the material by means of a milling cutter after the holes for 
the bolts have been drilled. At D is shown the so-called marine 
type of connecting rod head. In this design the rod proper is 
forged or cast with a T head, which is milled off to be fitted to 
the bronze crank pin box, which is cast in halves. This design 
is little used in automobile motors, but is extensively employed 
on marine motors, the rod being hand forged. 

In the head shown at E the cap is made of less width than the 
other half, the idea being that weight can be saved in this way. 



since under ordinary conditions of operation the maximum 
pressure on the cap is much less than on the other half of the 
head. It is to be understood that the half bushing in the cap is 
also shorter than the other one. The same idea underlies design 
F, in which a portion at the middle of the cap, between the two 
sets of lugs is cut away. Designs A, B and C are commonly 
used in automobile practice, while D, E and F are rare. 

Crank Pin Bearings—The bushings for the connecting rod 
head are made in halves. They are cast in metal dies of some 
soft bearing or anti-friction metal (babbitt or white bronze). A 

































































































THE CONNECTING ROD. 


211 


softer metal is employed than for the piston pin bushing, because' 
the unit pressure on the crank pin bearing is much smaller, and, 
besides, in the case of the crank pin it is essential that the wear 
be confined to the bushing as far as possible, because the renewal 
of a crankshaft means great expense, whereas a piston pin can 
be replaced cheaply. The bushings are often cast with integral 
pins or dowels which fit into drill holes on the inside of the con¬ 
necting rod head, and are cast or cut with oil holes and with oil 
grooves, the latter usually of spider-foot form. Instead of spider- 
foot grooves, some manufacturers cut intersecting longitudinal 
and circumferential grooves in the head. At the ends the bush¬ 
ings have external flanges equal in width and height to their 
thickness. 

Size of Studs —Ordinarily the connecting rod cap is held in 
place by either two or four studs and nuts. The necessary diam¬ 
eter of these studs can be calculated as follows: The greatest 
strain will come upon the studs when the piston starts on the 
auction stroke while the motor runs at “racing” speed, say, 2,400 
revolutions per minute in the case of our 4x5 inch motor. The 
force of acceleration then is 70.3 pounds per square inch of piston 
head area at 1,200 r. p. m., as we found in the chapter on the 
Conversion of Reciprocating Into Rotary Motion. At 2,400 revo¬ 
lutions it will be exactly four times as great, since the inertia 
force increases with the square of the speed. The total inertia 
force, then, is 

12.56 x 4 x 70.3 = 3,532 pounds. 

In the calculation of this force only one-half the connecting rod 
weight is taken into account, the other half being considered to 
have a rotary motion. This half, however, is subjected to cen¬ 
trifugal force, and at the moment when the inertia force on the 
reciprocating parts is a maximum the centrifugal force on the 
rotating part of the connecting rod is in the same direction as 
the inertia force, and should, therefore, be added to it. The 
weight of the connecting rod may be assumed to be 5 pounds, so 
that one-half of it is 2.5 pounds. The radius of rotation of this 
weight at the moment when the crank is in the top dead centre 
position is about 3 inches. Consequently the centrifugal force 
on this weight at 2,400 revolutions per minute would be (equation 
3i) 

1.226 X 3 X 2 X — = 1,47* founds. 

\ 60 / 12 

The total force on the cap bolts, therefore, is 

3,532 + 1,471 = 5,003 pounds. 

Figuring on a load of 22,000 pounds per square inch, the 



212 


THE CONNECTING ROD. 


necessary section of metal in each of two bolts is 

5,000 . , 

-—-=0.113 square inch. 

2 X 22,000 

The size of screw with S. A. E. thread whose cross-section 
at the bottom of the thread is nearest to this is fa inch. In 
case four bolts or studs are to be used, each must have a section 
of 0.057 square inch at the bottom of the thread. A fa inch 
screw with S. A. E. thread has nearly this section, but a 
inch stud would probably be used owing to the danger of 
injuring the smaller thread by a too powerful application of 
the wrench. 

Offset Connecting Rods—With the object of producing ? 
motor of the greatest possible compactness, and yet providing 
liberal bearing surfaces, many American designers offset their 
connecting rods, as shown on an exaggerated scale in Fig. 120. 
The following consideration shows that nothing is gained by 
this oroceeding. The resultant of the distributed pressure on 
the piston pin bearing must be equal to and in line with the 
esultant of the pressure at the crank pin bearing. Let us con- 


c: 



\ 


\ 





\ 


R 


sider the two bearings 
equally offset from the 
rod in opposite directions. 
Then the resultants of the 
pressures at die bearing 
surfaces will be repre¬ 
sented by R. These forces, 
it will be noted, are to 
one side of the centres of 
the bearings, and we found 
in the discussion of the 
side thrust of pistons that 
when a bearing surface 
is thus unsymmetrically 
loaded the pressure is un¬ 
evenly distributed over 
the surface. The unit 
pressure at the end far- 


~>-i thest away from the point 


Fig. 120. 


where the resultant takes 
effect is to the unit pres¬ 
sure at the opposite end 
as the distance of the 
nearest end from the re¬ 
sultant is to the distance 
of the farther end. Sup- 














































THE CONNECTING ROD. 


213: 


pose, for instance, that in Fig. 120 the two bearings are of equal 
length, and that the two “overhanging” portions of the bearings 
are equal to one-third the length of the bearing each; then the lo¬ 
cation of the resultant will be one-third the length from one end 
of the bearing and the unit pressure will be twice as great 
at the nearer end as at the farther end. Calling the aver¬ 
age unit pressure p, the unit pressure at the nearer end will 


be 


At 

3 


The wear at the near end will, of course, be pro¬ 


portional to —, or 33 per cent, greater than if the bearings 

were of the same size and in line. Now, if the overhanging 

•3 p 

portions were left off the unit pressure would be uniform, — 

throughout the length of the bearing, or only about 13 per cent, 
greater than with this big offset. In an offset rod the bearings 
will wear more rapidly at 
one end than at the other, 
and will then subject the 
connecting rod to addi¬ 
tional bending strain, and 
cause side thrust on the 
crankshaft and cylinder 
wall. 

The connecting rod cap 
is generally fitted with a 
“spoon” or quill, which 
dips into the oil at the 
bottom of the crank case 
once during every revolu¬ 
tion and projects it over 
the inner wall of the 
crank case, cylinder and 
piston (Fig. 121). If no 
such dipper were provided 
it would be necessary to 
carry the oil level at such 
a height that the cap it¬ 
self would dip into the 
oil, and this, at high mo¬ 
tor speeds, would be sure 
to splurge the oil entirely 
too much, causing the 
motor to smoke and the 
oil to be wasted. When 



















































214 


THE CONNECTING ROD. 


it is considered that the cylinder of a 4x5 inch motor requires 
much less than a drop of oil per revolution, it will be realized 
that the part dipping into the oil can be made very small. 

When splash lubrication is depended on for the rod, two slant¬ 
ing holes are drilled through the top half of the head, as shown 
in Fig. 121, through which any oil running down the connecting 
rod sides may reach the crank pin bearing surface. 

iManufacture of Rods—In the machining of the rods the 
greatest care must be taken to get the holes at the two ends 
parallel with each other. On a rod of the design herewith illus¬ 
trated there are five successive machining operations, and a set 
of jigs and fixtures for performing these operations was de¬ 
scribed by Robert G. Pilkington in The Horseless Age of 
November 25, 1908, whose drawings are herewith reproduced. 
The five successive operations are as follows: 

1. Milling sides of ends (one fixture). 

2. Drilling and reaming crank and piston pin fits (in jigs). 

3. Straddle milling bolt and nut surfaces on head (one fixture). 

4. Drilling and reaming small holes (one jig). 

5. Cutting off cap (No. 3 fixture). 




Fig. 122. 































































































THE CONNECTING ROD. 


215 


Fig. 122 shows a fixture for holding the rods while the sides of 
the ends are being milled off. It is arranged to carry eight pieces. 
In filling the fixture the first time after setting up, pieces are put 
in the odd numbered positions and a ‘‘dummy” in the even num¬ 
bers. The fixture is placed in a milling machine with two ad¬ 
justable vertical spindles, which are set to take off the required 
amount, and a cut is taken. Before reversing the feed, No. r 
rod is turned over and put in No. 2 position, No. 3 in No. 4, and 
so on. The hardened pads in the even numbered positions are 
just enough higher than those in the odd numbered ones to leave 
the work of the required size. After clamping down and running 
under the cutters a second time the rods on the even numbered 
positions are finished. Odd numbered positions are again filled 
with rough work. All pins and pads which locate the work are 
of hardened tool steel. The clamps are made to bear on the web 
of the rod, not on the flanges. 

A multiple spindle drill of at least four spindles is required 
for drilling and reaming the crank and piston pin holes. Fig. 123 
shows a box jig for this purpose. The usual slip bushings for 
drill and reamer are provided for each end. In putting the work 
into the jig it is laid on the lower bushings and the lid is brought 



Fig. 123. 
























































































216 


THE CONNECTING ROD. 


L 


H) [a 

E30 1 


Co)) «= 

©U O 

= i 

cL 


3 



Fig. 124. 


down and clamped by the cam hooks, one of which is shown at 
the right. The holding screws in the lid are then brought down 
lightly, followed by bringing up tight on the locating screw at 
the right. This forces the piston end of the rod into its V block 
and the crank end over against the 45 degree surface of its block. 
The work-holding screws are then brought down tight. 

The jig in Fig. 124 locates two holes for the cap bolts, an oil 
hole in each end and four dowel holes for the babbitt liners. 
The large locating pins are made hollow, which makes it very 
easy to clean all chips and lubricant from the jig with the air jet. 
The two upper dowel holes in the head are located positively by 
the holes in the locating pin serving instead of drill bushings. 



Fig. 125. 
























































































































































THE CONNECTING ROD. 


217 


The holding screw with the T head swings on the pivot indi¬ 
cated by the nut, and allows the work to be lifted out. 

The fixture shown in Fig. 125 serves to hold the rods for 
operations 3 and 5. None of the pressure of the cut is intended 
to be taken by the spring pressed live centre, the side pressure 
being neutralized by the use of opposing cutters. In setting up 
the work four pieces are placed on the arbor, the cap is put on 
the end and the key dropped in place. After placing on the cen¬ 
tres the locating pin is pushed through the lower ends and the 
key in the arbor cap is tapped in place, clamping the work firmly. 
Exactly the same procedure is carried on in the fifth operation,, 
with the difference that thin slitting cutters are used. 


CHAPTER XI. 


VALVES AND VALVE GEARS. 

The modern standard automobile motor works on the four 
•cycle principle, as explained in the chapter on “The Otto Cycle,” 
and is fitted with mechanically operated poppet valves for both 
the inlet and the exhaust. A poppet valve consists of a disc of 
metal, with a stem on one side coaxial with the disc, which 
closes a circular opening in a partition wall between two cham¬ 
bers, against which wall it is drawn by a spring. To open the 
valve a force has to be applied to it contrary to the direction 

of spring pressure, strong enough to 
overcome the spring. Formerly the in¬ 
let valves were generally actuated by 
the suction prevailing in the cylinder 
during the inlet stroke, such valves be¬ 
ing called automatic or suction valves. 
A typical valve of this kind with its 
cage is illustrated in Fig. 126. It will 
be understood that, since the valve opens 
toward the cylinder, when the piston 
moves down on the inlet stroke and the 
pressure in the cylinder drops below at¬ 
mospheric, the excess pressure of the 
atmosphere against the outer face of 
the valve head overcomes the pressure of the light valve spring, 
and the valve opens. Toward the end of the inlet stroke the 
•excess pressure of the atmosphere diminishes and the spring then 
closes the valve. 

Since automatic inlet valves are no longer used on automobile 
motors, except in rare instances, it is not necessary to go ex¬ 
tensively into their theory, but it may be well to point out some 
•of the disadvantages which led to their abandonment in favor 
of mechanically operated valves. When the motor is pulled 
•down in speed by a heavy load the automatic valve will close as 



Fig. 126.—Automatic 
Inlet Valve With 
Cage. 


218 


































VALVES AND VALVE GEARS. 


219 


soon as the suction force decreases to the point where it is 
equaled by the force of the valve spring; that is, before the end 
of the suction stroke, and the cylinder receives an incomplete 
charge."' The valve will not remain open under less than a cer¬ 
tain suction, and the motor, therefore, cannot be throttled down 
as much as a motor with mechanically operated valves. In 
other words, it is less flexible. The automatic inlet valve has a 
tendency to stick to its seat, owing to the gumming effect of the 
gasoline, and to produce noise by “fluttering” and by “wire¬ 
drawing” the charge. 

Angle of Seat—The poppet valve consists of two parts, the 
head and the stem. That portion of the head which when the 
valve is closed rests on a machined portion of the cylinder cast¬ 
ing is known as the seat. This surface usually forms a truncated 
cone whose generatrices make an angle of 45 degrees with the 
plane of the valve head. Sometimes, however, the valve seat is 
a plane surface perpendicular to the valve axis, and, again, the 
seat is a truncated conical surface of less than 45 degrees in¬ 
clination. Flat seated valves are used by the Daimler Motor 
Company, of Cannstadt, Germany, in their Mercedes motors, b> 
the Knox Automobile Company, of Springfield, Mass., and 
by a few others, but the great majority of makers use 45 degree 
conical valves. The advantage of the flat seated valve is that for 
a certain area of opening at the seat it does not need to be lifted 
as high as a conical valve. It can, therefore, attain its maxi¬ 
mum opening quicker and retain it longer, and since the amount 
of travel in closing is less,- it should operate more quietly. The 
conical valve, on the other hand, has the advantage that it is self¬ 
centering. The two types are shown in Figs. 127 and 128, 
respectively. 

In Fig. 127 let cf ( = B B') be the clear diameter of the valve or 
the diameter of the bore of the valve seat, and h (=BD) be the 
lift. Also, let 6 be the angle of the valve seat. It is required 
to find the effective area of opening corresponding to any lift 
h. Referring to the figure, the minimum area of cross-section is 
evidently a truncated conical surface, of which B C is a genera¬ 
trix. Now, 

B C = B D cos C B D = h cos d. 

Also 

C £ = B C cos B C E = 

B C sin C B E = 
h cos 0 sin 0 . 

Hence the diameter C C r is 


220 


VALVES AND VALVE GEARS. 




Fig. 127. 


Fig. 12 C 


d -f- 2 h cos 0sin 6 . 

The area of the truncated conical surface is 


7r 


CC' -f- B B' 
2 


XBC = 


ir (d -j- h cos 6 sin 6 ) h cos 6 — 

tt {d h cos 6 h 7 cos 2 0sin 6 ) . (87) 

In case 0 = 45 degrees the expression for the area becomes 
// - 7r (0.707 d h + 0.353 /1 2 ) . (88) 


It will readily be seen by reference to Fig. 128 that in the case 
of a flat seated valve the area of opening is 7 rd h. Hence, calling 
the lift of the flat valve In and that of the 45 degree valve h « 
the two lifts must be so related that 


0.707 tf,* 45 + 0.353 /! 2 45 = dhu 
from which it follows that 

hi = o. 707 £45 + o.353 ~-‘ 

d 

If the lift of the 45 degree valve is one-fifth the clear diameter, 
then 


hi = o. 778 

or roughly four-fifths of h i6 . 

It would be desirable in one respect to have the passage around 
the valve when fully lifted about equal to the area of the valve 
port. The full area of the valve port is 0.7854 d 2 , but the stem 
will reduce this to 0.74 d 2 . Hence 

3.14 (0.707 dh + 0.353 h 2 ) =0.74 d 2 , 

2.22 d h + 1.11 h 2 = 0.74 d 2 , 
which when solved gives 

h = 0.29 d. 

In practice, however, the lift is always made less than this, 
owing to the difficulty of making the valve action sufficiently quiet 
with high lifts. There is also a noticeable tendency to make the 
lift of small valves proportionately greater than that of large 
valves. A good rule is to make the lift 

A = % inch.(89I 

O 







































VALVES AND VALVE GEARS. 


221 


Material of Valves —The heads of the valves are subjected 
to the high temperature of the burning gases, and it is essential 
that they should not warp under the influence of the heat, and 
that their seats should not burn or oxidize, as in either case they 
will b ecom e leaky. Pure nickel is sometimes used on account of 
its non-oxidizing qualities. A more common’ material for valve 
heads is high percentage (30-36 per cent.) nickel steel, which has 
an exceedingly low co-efficient of heat expansion, hence does not 
warp. Recently tungsten steel (high speed tool steel) has come 
into extensive use, and valves of it are said to be non-pitting, 
because of their extreme hardness. The stems of the valves are 
preferably made of carbon steel, which is harder than the high 
percentage nickel steel, and therefore stands the shocks better, and 
it is also considerably cheaper. Poppet valves with high per¬ 
centage nickel steel heads electrically welded to carbon steel 
stems are at present very widely used. Valves with cast iron 
heads screwed and riveted to carbon steel stems are also used to 
some extent. The bottom ends of the stems should be case 
hardened (which is conveniently done in cyanide of potassium), 
so that they will not wear or become deformed under the shocks 
of the push rods. 

^Form and Dimensions of Head —The valve head should be 
made of the minimum thickness that will insure against deforma¬ 
tion, and to this end it is sometimes made dome shaped. In 
American practice the head is made flat on the under side and 
rounded on top, so as to give increased strength at the centre, 
where the bending moment due to the explosion is the greatest. 
The principal dimension of the valve head is the bottom diameter 
d, which should be equal to the clear diameter of .the valve port. 
All the other dimensions of the head may be expressed in terms 
of d as follows (for 45 degree valves) : 

Outside diameter of head = 1.15 d. 

Thickness at centre of head = 0.15 d. 

Valve stem diameter = 0.15 rf + 0.15 inch. 

Radius of fillet at junction of head and stem = 0.2 d. 

The above ratio between the bottom and top diameters of the 
valve head gives a seat on the head about o. 10 d in width, and 
the projection of this width on the horizontal plane is 0.07 d. The 
latter should be the same for valves with other seat angles. The 
width of the seat formed in the cylinder casting must be some¬ 
what less, so there will be no interference between the outer edge 
of the valve and the seat in the casting after repeated grindings. 
European designers sometimes provide an exceedingly large fillet 
between the valve head and the stem, of a radius equal to and 
even greater than the overhang of the head, to deflect the col- 


222 


VALVES AND VALVE GEARS. 


umn of gas and thus lessen the resistance of the valve passage, 
and make the upper surface of the head flat, as shown in Fig. 
129. They also often make the upper portion of the stem slightly 
less in diameter than that portion working in the guide, so that 
there is no possibility of a shoulder being formed where the stem 
leaves the guide when the valve is in the closed position. This 
is also shown in Fig. 129. 

If the head is made of cast iron it is provided with a sort of 
hub on the under side and is screwed over the reduced upper end 
of the stem against a shoulder, and the end of the stem is then 
riveted over, as shown in Fig. 130. All valves must have a screw¬ 
driver slot milled in the head into which the grinding tool may 
engage. 

Mean Gas Speed Through Valves—In calculating the di¬ 
mensions of valves, the calculations are often based upon a 
certain mean gas speed through the valve port in feet per 
minute. It is obvious that the mean gas speed through the 



Fig. 129.—European Design Fig. 130.—Cast Iron Head 
Poppet Valve. Poppet Valve. 



valve port is to the mean piston speed as the piston head 
to the valve port area. But since the piston head area is pro¬ 
portional to the square of the bore and the valve port area 
to the square of the clear valve diameter, the mean gas veloc¬ 
ity through the valve port is to the mean piston speed as the 
square of the bore is to the square of the clear valve diam¬ 
eter— 


Vy : v v = b 7 : d 2 , 

Hence 



In modern engines v y has a value of 7,200 to 8,000 feet per 
minute when the engine runs at its speed of maximum output. 

A committee appointed in 1910 by the (British) Institution 
of Automobile Engineers to develop a rating formula for 





















VALVES AND VALVE GEARS. 


225 

gasoline motors, in its report suggests the following formula 
for the clear diameter of the inlet valve: 


d = o. 2 gb 




where b and l have their usual meanings and m is the weight 
of the reciprocating parts (piston and connecting rod). The 
idea underlying this recommendation* is that a motor will be 
able to run at a higher piston speed the greater the length of 
stroke l and the less the weight m of the reciprocating parts, 
hence the valve diameter should increase with the former 
and as the latter decreases. It is admitted, however, that 
this formula in the majority of cases gives valve diameters 
greater than actually employed in practice. 

Degree of Filling—A gasoline engine cylinder in normal 
operation never draws as much charge per suction stroke as 
is represented by the piston displacement volume at atmos¬ 
pheric pressure and temperature. This latter amount is re¬ 
ferred to as a complete charge. The ratio of the amount of 
air actually taken in per suction stroke to a complete charge 
is known as the volumetric efficiency. This factor decreases 
as the speed of the motor increases, increases and decreases 
with the size of the inlet valves, and varies also with the 
valve timing and the cylinder wall temperature. Some tests 
made on a Franklin four cylinder, 4x4 inch air cooled motor 
at Cornell University, and reported in a paper read by Prof.. 
R. C. Carpenter before the Society of Automobile Engineers 
at New York in January, 1911, showed that this motor, fitted 
with poppet valves of a clear diameter of i*/s inches, opening 
5 degrees past dead centre and closing 30 degrees past dead 
centre, had a volumetric efficiency when cold (that is, when 
driven by an electric motor) of 82 per cent, at 1,000 r. p. m.. 
and 68 per cent, at 1,500 r. p. m. When run under its own 
power, with the inlet valves opening 5 degrees past dead 
centre and closing 30 degrees past dead centre, the volumtric 
efficiency was 68 per cent, with applied phosphor bronze cool¬ 
ing fins and 62 per cent, with cast iron integral cooling fins 
at 1,000 r. p. m.; 63 per cent, with applied phosphor bronze 
and 56 per cent, with integral cast iron cooling fins at 1,500 
revolutions per minute. This clearly shows the effect of the 
cylinder wall temperature on the volumetric efficiency, the 
applied phosphor bronze fins evidently cooling less than the 
integral cast iron fins. The effect of valve timing on the volu¬ 
metric efficiency is shown by the following figures: With cast 
iron integral fins but with the inlet valves opening 7 degrees 


224 


VALVES AND VALVE GEARS. 


before dead centre and closing 17 degrees past dead centre the 
volumetric efficiency was 45 per cent, at 1,000 revolutions and 
47 per cent, at 1,500 revolutions. Curves plotted to show the 
variation of the volumetric efficiency with the speed of the 
engine bring out the interesting fact that whereas in the cold 
engine the volumetric efficiency decreases practically uni¬ 
formly as the engine speed increases, in the hot engine it de¬ 
creases rapidly at first, then less rapidly and becomes constant 
or begins to increase at from 1,300 to 1,400 revolutions per 
minute. The probable explanation of this difference in the 
variation of the volumetric efficiencies of hot and cold engines 
is that at high rotative speeds the incoming gases do not have 
sufficient time to absorb much heat from the cylinder wall, 
hence the difference in the efficiencies of hot and cold cylin¬ 
ders is then less. 

Valve Guides—Sometimes the guides for the valve stems are 
cast integral with the cylinders, but more frequently they are 
inserted. Casting them integral is the least expensive construc¬ 
tion, but it is open to the objection that owing to the unavoidable 
inaccuracies in molding, the holes through the guides are fre¬ 
quently considerably out of centre, and, besides, when the guide 
has become worn it is very difficult to refit it so as to make a 
fairly tight joint, which is essential particularly in connection 
with the inlet valve. Since it is impossible to properly lubricate 
the exhaust valve guides, wear on these is comparatively rapid. 

The methods of inserting the guides differ considerably. For¬ 
merly it was customary to provide them with a screw thread and 
screw them into the cylinder casting, as shown at A, Fig. 131. 
This is objectionable, however, since it is difficult to get the thread 
on the guide absolutely concentric with the hole through it, and 
if the guide should ever be removed, and in replacing not be 
turned to exactly the same angle, the hole through it would not 
be concentric with the valve seat. It is therefore preferable to 
clamp the guide into the cylinder casting by means of a flange on 
it and a nut, as shown at B, or to force it into the casting under 
pressure and hold it in position by one or two flat head machine 
screws, as shown at D. The screws can even be dispensed with, 
since the valve spring pressing against the flange of the guide will 
hold it in its place in the cylinder casting, but the screws will 
prevent it from turning. Instead of inserting the guides from 
below, they may be inserted into the cylinder casting from above, 
as shown at C. In this case the valve spring bears against a seat 
formed on the cylinder casting, and the guide is held in place by 
friction and by its weight. 


VALVES AND VALVE GEARS. 


225 



Fig. 13 i.—Valve Guides. 


The mean diameter of the guides is made about twice the diam¬ 
eter of the valve stems, and the length is made 1 to 3 times 
the clear diameter of the valve, the former figure being used 
where overhead valves or similar constructions limit the available 
space. 

Valve Timing—The valves are operated by cams on a shaft 
which turns at one-half the speed of the crankshaft, so that each 
valve is opened and closed once during two revolutions of the 
crankshaft. For operation at very low speed the inlet valve 
should begin to open and should close at the beginning and end 
of the inlet stroke, respectively, and the exhaust valve should 
open a trifle before the end of the power stroke and close at the 
end of the exhaust stroke. However, in order that the motor may 
operate satisfactorily at high speed, it is necessary to alter the 
valve periods considerably. The exhaust valve must open earlier 
and close later, and the inlet valve must open later (in order 
not to overlap the exhaust opening) and close later. The re¬ 
spective lags and leads should be greater in proportion as the 
motor is intended to run at higher rotary speed. For a motor in¬ 
tended for service at normal speed, say a 4x5 inch motor to run 
at 1,200 revolutions per minute, the following valve timing would 
be suitable: 

Exhaust opens 40 degrees ahead of bottom dead centre. 

Exhaust closes 5 degrees past top dead centre. 

Inlet opens 10 degrees past top dead centre. 

Inlet closes 20 degrees past bottom dead centre. 

In motors intended to run at very high speed, and consequently 
provided with valves of very large diameters, the timing may be 
made as follows: 

Exhaust opens 45 degrees ahead of bottom dead centre. 





































226 


VALVES AND VALVE GEARS. 


Exhaust closes io degrees past top dead centre. 

Inlet opens 15 degrees past top dead centre. 

Inlet closes 30 degrees past bottom dead centre. 

The last mentioned timing system is illustrated diagrammatically 
in Fig. 132. 

It should be pointed out that if a motor is timed to give the 
very best output at high speeds it will not run so satisfactorily at 
low speed, and will not be as flexible. This is due to the fact that 
when the exhaust valve opens very early, some of the power other¬ 
wise available at low speed is lost through the exhaust, and when 
the inlet valve closes very late some of the charge drawn in dur- 



Fig. 132.—Valve Timing Diagram. 


ing the suction stroke will be forced out again during the begin¬ 
ning of the compression stroke. The cylinder, therefore, will not 
receive a full charge, and will not utilize to the fullest advantage 
the charge it does receive. 

Influence of Inlet Valve Timing on Power and Fuel 
Efficiency—Some tests were made in 1909 at the Laboratory 
of the Automobile Club of France on a four cylinder Renault 
motor of 100 mm. bore and 140 mm. stroke, to determine the 
effect of variations in inlet valve timing on the power output, 
fuel efficiency and torque. The motor was provided with 45 
degree poppet valves, inlet and exhaust valves being identical. 
The valve heads had a bottom diameter of 36 mm., a top 
diameter of 42 mm. and were lifted 7.5 mm. The inlet 














VALVES AND VALVE GEARS. 


227 


valves began to open 15 degrees after the top dead centre, 
and by means of a stepped cam could be closed 27 degrees 
ahead of the lower dead centre, at the dead centre, and 19, 27, 
40, 60 and 90 degrees past dead centre. The exhaust valves 
began to open 33 degrees ahead of dead centre and closed on 
dead centre. The effects of these changes in the timing of 
the inlet valve are graphically shown in Figs. 133 and 134, 
and the results of the tests may be summarized as follows: 

As the inlet valve is set to close later, the power for a 
given speed of rotation increases until a lag of 60 degrees is 



Fig. 133.—Effect of Inlet Valve Timing on Power Output. 

reached. When the lag of the inlet valve is still further in¬ 
creased, to 90 degrees, the power diminishes, but the diminu¬ 
tion in power becomes less as the speed of the motor in¬ 
creases, and for very high angular speeds such a lag would 
seem to give an increase in power. The angular speed corre¬ 
sponding to the maximum power is the greater the later the 
inlet valve closes. 

The effect on the fuel economy is practically the same as 
that on the power output. That is, the amount of fuel con¬ 
sumed per horse power hour decreases as the inlet valve is 



















228 


VALVES AND VALVE GEARS. 


closed later, up to a lag of 60 degrees, but increases for a 
lag of 90 degrees. 

It will be observed that the tests covered the operation of 
the motor at high speed only, from about 1,000 feet per min¬ 
ute piston speed upward. It is obvious that inlet valves with 
as big a lag as 60 degrees would not be very satisfactory for 
operation at low speed, and for this reason on stock motors 
the lag is always made considerably less than this. But the 
test results indicate that in racing motors an inlet valve lag 
of 60 degrees or more would be advantageous. 

The reason the inlet valve has to close some time after 
the completion of the inlet stroke is that at high speed the 
inrushing column of gas possesses considerable inertia, anc 
the charge continues to flow into the cylinder even after the 
piston has started on the return or compression stroke. 



Fig. 134.—Effect of Inlet Valve Timing on Fuel Efficiency. 

Cams and Cam Followers—The rate at which the valve 
is opened and closed depends upon the cam outline and upon the 
type and size of cam follower employed. From the standpoint 
of gas flow it is, of course, desirable that the valve should open 
and close very quickly, and remain fully open for the greatest 
possible length of time. However, the valve gear must operate 
quietly, and in order to do this it must lift and drop the valves 
more or less gradually. 

There are three common types of cam followers. The most 
extensively used is the cam roller; next comes the mushroom type, 
and Anally the rounded or V type. The general method of lav- 
















VALVES AND VALVE GEARS. 


229 


ing out a cam for use with a roller follower is as follows: Sup¬ 
pose that the cam is to operate the exhaust valve, which is to open 
40 degrees ahead of the bottom dead centre and close 6 degrees 
beyond the top dead centre. The valve must then remain open 
for a period corresponding to 

40 + 180 + 6 = 226 degrees of crank motion — 

113 degrees of camshaft motion. 

Referring to Fig. 135, we first lay out a circle known as the 
base circle, which in this case can be made 1% inches in diameter. 
The diameter of this circle should be from three to four times 
the valve lift, the higher ratios being used in low speed motors. 
The cam roller may be 1 inch in diameter and the maximum lift 
of the valve % inch. Next, we draw two radii making an angle 
of 113 degrees with each other, and then a circle concentric with 
the base circle but %2 inch larger in diameter, known as the 



clearance circle, which allows for a clearance of 1-64 inch be¬ 
tween the valve and the push rod or between the cam follower 
and the cam. Next, taking centres on the radial lines O A and 
O B, we strike circles equal in diameter to the cam rollors and 
tangent to the clearance circle. Then we draw straight lines C D 
and E F tangent to both the roller circles and the base circle. 
Next we draw an arc of a circle G H concentric with the base 
circle and of a radius equal to the radius of the base circle plus 
the lift of the valve—in this case 1 inch. After rounding off 
the corners at the intersections of the full lift circle and the 
tangents we have the complete cam outline. This form of cam 
is known as a tangential cam. 










230 


VALVES AND VALVE GEARS. 


Cam Opening Curve—In the United States the tangential 
cam is used almost exclusively, but it is not entirely satisfactory, 
owing to the fact that it begins to lift the valve and allows it to 
strike its seat in closing rather abruptly, producing severe strains 
in the valve and causing considerable noise. The valve should 
evidently begin to lift as gradually as possible and its motion 
should be reduced practically to zero before it reaches its seat. 
The most advantageous outline of cam from this point of view 
can be arrived at by means of the following reasoning: To lif* 
the valve force must be applied to it. In order to lift it a certain 
distance in a certain time with minimum force, the force 
should be constant throughout its period of action. At tie 
moment the valve attains its full lift its velocity is zero; tie 
force applied to it accelerates it to its maximum velocity, and 
after the force is withdrawn the valve must be allowed a certa n 
time to decelerate. Since the total acceleration and deceleration 
are equal, it seems logical that the same amount of time should 
be allowed for each. If, then, we allow a certain angular motion 
during which the valve must attain its maximum lift, it must 
acquire its maximum upward velocity during one-half this mo¬ 
tion. With the tangential type of cam shown in Fig. 135 it re¬ 
quires an angular motion of 42^2 degrees of the cam shaft to 
fully lift the valve, and we will allow the same in this case. It 
may be pointed out that with a tangential cam the angle of lift 
and drop, if we neglect the rounding at the intersection of the 
tangents and the full lift circle, is determined by the ratio of 
the base circle radius to the lift. With a certain lift the greater 
the diameter of the base circle the smaller will be the angle of 
the lift and drop; that is, the faster the valve will be opened 
and closed. 

Cams designed to give a constant acceleration to the valve and 
allow it to decelerate at the same uniform rate are known as 
constant acceleration cams, and were first suggested by the writer 
in 1905 (see The Horseless Age of September 20, 1905). They 
have recently come into extensive use abroad. The characteristic 
feature of these cams is that their lifting and dropping sections 
are concave, instead of being straight, as in the tangential curve. 

Laying Out Constant Acceleration Cam—A constant ac¬ 
celeration cam may be drawn as follows: Suppose that the base 
circle is of 1% inches diameter, the cam roller 1 inch, that, the 
lift is Vz inch and that the valve is completely lifted and dropped 
in 42.5 degrees of cam motion. The valve timing is supposed 
to be the same as in the previous example, the valve remaining 
open for 113 degrees of cam motion. 


VALVES AND VALVE GEARS. 


231 


In the diagram Fig. 136 this cam motion of 113 degrees is 
laid off along the axis of abscissas, and momentary valve posi¬ 
tions are plotted along the axis of ordinates. At 42.5 degrees 
from the point where the valve begins to open it attains its full 
lift of inch. At one-half this angular distance it attains one- 
half its full lift, or I® inch, since the valve accelerates uniformly 
during the first half of the lifting period and decelerates at the 
same rate during the second half. In order to find the lift 
corresponding to intermediate angular positions of the cam, say, 
5, 10, 15 and 20 degrees from the beginning of the lift and the 
end of the lift, respectively, we take into account the fact that 
whenever a body is uniformly accelerated the distance passed 
through by it in any particular time is proportional to the square 
of that time. In the present case the time is proportional to the 
crank motion, and hence the lift will be proportional to the square 
of the crank motion— 

h—ct\ 



Fig. 136. —Diagram Showing Lift and Drop of Valve by 
Constant Acceleration Cam. 

Since for an angular motion of 21.25 degrees the lift is i *5 inch 
the value of the constant c in the above equation is 

3 

- .. T? =0.000415, 

21.25^ 

and the lifts at various angular distances from the point at which 
the valve begins to lift are as follows: 

5 0 — 0.000415 x 5 2 = 0.0104 inch. 
io° — 0.000415 x io 2 = 0.0415 inch. 

15 0 —0.000415 x i5 2 = 0.0934 ^ch. 

20°—0.000415 X20 2 = 0.166 inch. 

These four points enable us to draw the first half of the lifting 
incline. The second half is a duplicate of this, but reversed. The 
dropping incline is identical with the lifting incline and can, 
therefore, be drawn-with the aid of the same data. 

After having laid out the valve lifting and dropping diagram, 










232 


VALVES AND VALVE GEARS. 


Fig. 136, the cam itself is drawn as follows: Lay off the base 
circle and the clearance circle, and then draw two radial lines 
from the centre of these circles 113 degrees apart. These radii 
evidently denote the opening and closing positions of the cam. 
Next draw radial lines between the two already drawn, one at 
42^ degrees from each. These denote the positions of the cam 
when the valve attains its full lift and when it begins to drop. 
Next lay off radial lines 5, 10, 15 and 20 degrees from each of 
the four lines already drawn, as shown in Fig. 137. Next strike 
a circle with a radius equal to the radius of the clearance circle 
plus the radius of the roller, intersecting all of the radial lines 
drawn. From this circle outwardly along the radial lines lay 
off distances equal to the lift of the valve corresponding to the 



Fig. 137.—Constant Acceleration Cam. 


particular angle, and with the joints so found as centres lay 
off circles equal in diameter to the cam roller. The curve tangent 
to all of these circles on the side toward the base circle is the 
cam outline. 

The lifting and dropping inclines thus obtained join the clear¬ 
ance circle, an imaginary line. In the cam proper they must 
join the base circle, and to this end the concave incline may be 
continued by convex circular arcs of a radius equal to two- 
thirds the base circle, with centres on the opening and closing 
radial lines, and tangents drawn to these arcs and the base circle. 
For best results the valve gear must be adjusted to the exact 
clearance intended, 1-64 inch. 

Owing to the fact that these cams require a rather heavy spring 
it is a good plan to make the acceleration during lift and the 
deceleration during closing different from the deceleration during 
lift and acceleration during closing. For instance, in the above 







VALVES AND VALVE GEARS. 


233 


case the acceleration during lift might be completed in 17.5 
degrees and the deceleration in 25 degrees. 

Mushroom Type Cam Followers—The mushroom type of 
cam follower is slightly simpler in construction than the 
roller cam follower, and it also reduces the number of wearing 
parts by eliminating the roller pin. But since it substitutes sliding 
motion for rolling motion at the cam surface it is doubtful 
whether it reduces the wear in the valve gear on the whole. 

The method of laying out a cam for a mushroom type follower 
is as follows (Fig. 138) : First draw the base circle and the top 
circle, the radii of which must be assumed. The two centres 
must be placed a distance 

D — R c l — r 

apart. Now lay off the clearance circle C and a tangent T to 
this circle at a point P distant from the point P' where the line 



Fig. 138.—Exhaust Cam for Mushroom Type Cam Follower. 

connecting the centres of the base and top circles cuts the clear¬ 
ance circle, an angle equal to one-half the valve period angle a. 
Then draw an arc A which is tangent to the base circle, the top 
circle and the tangent T to the clearance circle. The centre of 
this arc is found to be located at O. This arc forms the flank 
of the cam protuberance. A peculiarity of the mushroom type 
cam and follower is that during the lift the valve is accelerated 
very strongly and then decelerates at a much slower rate. This 
permits of using a comparatively weak spring. Speed and accel¬ 
eration diagrams can be constructed graphically by working on' 
a large scale. An analytical solution of the speed and accelera¬ 
tion problem for this type of cam was published by Otto M. 
Burkhardt in The Horseless Age of January 28 and February 
4, 1914. 

















234 


VALVES AND VALVE GEARS. 


In drawing the valve opening diagram for this mechanism we 
proceed as follows: First radial lines are drawn from the centre 
of the base line at angular distances of, say, 5 degrees. Then 
tangents are drawn to the cam outline, so as to cut the radial lines 
at right angles. The distance of the points of intersection thus 
found from the clearance circle represents the lift of the valve 
corresponding to the angular position of the cam represented by 
the particular radial line. With this type of cam follower the 
valve does not reach its maximum lift until midway between the 
opening and closing positions, and immediately begins to close. 
The opening curve, as shown in Fig. 139, meets the base line at 
a rather sharp angle. 

Cams to be used in connection with rounded or V-shaped fol¬ 
lowers can be laid out by the same method as cams for roller 
followers. 



Fig. 139.—Valve Lifting Diagram With Mushroom Type 

Cam Follower. 


Comparison of Cam Outlines—The value of a cam mechan¬ 
ism for the operation of gasoline engine valves depends essen¬ 
tially upon three factors, as follows: 

1. Uniformity in the variation of the lifting speed. 

2. Value of the lift-angle integral. 

3. Average value of side thrust on cam follower. 

The speed of the valve is represented by the inclination of the 
valve opening curve. This curve should gradually merge into 
the base line, because if it meets the base line at a sharp angle 
the valve acquires a certain definite velocity in an instant, which 
means a shock and consequent noise. The greater the angle at the 
junction or intersection of the opening curve and the base line the 
greater the shock and the greater the noise and strain, other 
things being equal. 

The lift-angle integral, the second factor determining the value 
of the cam mechanism, may be explained as follows: The chief 































VALVES AND VALVE GEARS. 


235 


object in designing a valve oper¬ 
ating mechanism is to insure that 
the greatest possible quantity of 
gas will flow through the valve 
into or out of the cylinder during 
the period of valve opening. Now, 
the quantity of gas passing 
through an opening depends upon 
the area of opening and upon the 
time the flow continues. If the 
area of the opening constantly va¬ 
ries then the flow will depend upon 
the integral of the product of the 
instantaneous area of opening by 
the differential of the time. This 
integral is represented by the area 
enclosed by the valve opening dia¬ 
gram. Therefore, the greater the 
area enclosed by the valve open¬ 
ing diagram the better the cam 
form from this point of view. 

The above statement is based on 
the assumption that the area of 
valve opening is proportional to 
the lift, which is practically 
but not absolutely correct. 

The vertical reaction be¬ 
tween the cam surface and 
the cam follower is equal to 

the sum of the spring pressure and the inertia force. 
The latter is positive during the first half of the lift and 
the last half of the drop, and negative during the last half of the 
lift and the first half of the drop. However, the reaction is 
vertical only at no lift and at full lift. For all intermediate posi¬ 
tions the reaction is at an angle, as indicated in Fig. 140. Calling 
the vertical pressure P, it is evident that the pressure or reaction 
at the contact surface between cam and cam follower is 



Fig. 140.—Horizontal and 
Vertical Components of 
Reaction Between Cam 
and Roller. 


Pt = 


cos 0 


and the side thrust on the push rod guide 

P g =P tan 0. 

i 

Ratio of Base Circle Diameter to Lift —Considering only 
tangential cams, the rapidity of valve opening, the shock at the 















236 


VALVES AND VALVE GEARS. 


moment of opening and the side thrust on the push rod guide for 
various lifts depend upon the ratio r of the base circle diameter to 
the total lift. The limiting values for this ratio in ordinary prac¬ 
tice were given previously as 3 and 4. In Fig. 141 are shown two 
tangential cam outlines with base circle diameters equal to three 
and four times the total lifts, respectively, the cam roller diameter 
being three times the total lift. Disregarding the usual rounding 
at the junction of the tangents and the full lift circle, the valve 
attains its full lift in a time corresponding to a cam motion of 40 
degrees when r - 4 and 45^2 degrees when r =3. Thus the valve 
is lifted considerably faster when the base circle is relatively 
larger in diameter. In American practice the ratio r is generally 
equal to * 



Fig. 141.—Tangential Cams with Different Ratios of Base 

Circle Diameter to Lift. 


The cam having the smaller ratio of base circle diameter to 
total lift appears much steeper than the other one, and it is there¬ 
fore interesting to determine the values of angle 0 . In Fig. 142 
is plotted a curve for each of these cams, showing the angularity 
0 of the thrust for any angle of cam motion from the point of 
valve opening. It will be seen that the angularity at the begin¬ 
ning of valve opening is least with the smaller base circle, but 
as the lifting progresses the angularity with the small base circle 
becomes the greater. The maximum values are 

e ~ 2 > 2 ° 30' for r = 4, and 
0 = 34° for r = 3. 

These maximum angularities would, of course, be less if the 






VALVES AND VALVE GEARS. 


237 


corners of the cam were rounded off, as they always are in prac¬ 
tice, but .this rounding would reduce them in the same propor¬ 
tion. Another effect of the rounding would be that the angularity 
would vanish more gradually than shown in the diagram. The 
cam with the larger base circle will open and close the valve 
more quickly, will strike the cam follower a slightly sharper 
blow, and require a somewhat stronger spring. 

Offset Cams —Cams with their centre lines offset from the 
axis of the cam follower in the direction of the cam travel— 
that is, toward the right if the cam turns right handedly, and 



Fig. 142.—Angularity of Cam Thrust. 


vice versa—have been used to a certain extent in order to de¬ 
crease the side thrust on the push rod guide. The dropping slope 
of such cams must be very much convexed in order to allow the 
valve to close quietly. The advantage gained by using offset 
cams is due to the fact that the valve spring exerts a stronger 
pressure when the valve is fully lifted than when it is closed. 
The vertical reactions between the cam and push rod during the 
various phases of valve travel are as follows: 

First half of lift—spring pressure plus inertia. 

Second half of lift—spring pressure minus inertia. 























238 


VALVES AND VALVE GEARS. 


First half of drop—spring pressure minus inertia. 

Second half of drop—spring pressure plus inertia. 

Hence, if the spring pressure did not change during the lift, 
the average reaction between cam and cam follower would be 
the same during the lift as during the drop, and the conditions 
would be entirely unlike those in the crank and piston motion. 
As it is, the advantage to be realized is so slight that it is not 
worth the trouble involved in placing the camshaft off centre 
and giving the cams an irregular shape. 

Spring Pressure Required—The minimum pressure of the 
valve spring must be such that the exhaust valve will not open 
during the suction stroke when the throttle is nearly closed. 
Take, for instance, a valve with a clear diameter of I Y\ inches. 
The area of the under side of the valve then is 2.4 square inches. 
As the suction may attain 10 pounds per square inch, the mini¬ 
mum pressure of the spring should be at least 24 pounds. The 
mean pressure of the spring, that is, the pressure when the valve 
is half lifted, can be calculated as follows: 

The spring must hold the cam follower in contact with the 
cam, even at the maximum speeds at which the motor is likely 
to be run in service; say, one and one-half times the normal 
speed, or 1,500 feet piston speed per minute. For a 4x5 inch 
motor this would be 1,800 r. p. m. Suppose that the lifted parts 
—valve, cam follower and one-half of the spring—have a weight 
of 134 pounds. We found that the valve lifts 3 /$ inch while the 
cam turns through 42.5 degrees, and & inch while it turns 
through 21.25 degrees. During the time the cam makes this 
motion the valve accelerates and decelerates uniformly. The 
time required to make this motion at 900 revolutions per minute 
of the camshaft is 


60 X 21.2s 

— 0.004 secon 

900 X 360 

Since the acceleration is 


d. 


t 2 

we have for its value in this case 


1 


2 X 3/16 . 

-»- — 23,450 inches — 1,954 feet per second per second. 

0.004 2 

Now the force F required to impart to a weight W an accel¬ 
eration a is 


Hence, substituting values, 


Wa 

__ • 

g 


F = 


i-75 X 1,954 
32.16 


106 pounds. 





239 


VALVES AND VALVE GEARS. 


The mean pressure of the valve spring (its pressure when the 
valve is half lifted) should therefore be 106 pounds. The com¬ 
putation is based on the supposition that the acceleration and 
deceleration of the valve are uniform. With a tangential cam this 
is not the case, and the spring should then be made slightly 
stronger. By laying out the cam and follower on a large scale 
it is possible to determine the acceleration of the valve at differ¬ 
ent points of its lift, graphically, with a fair degree of accuracy, 
and the spring can then be calculated on the basis of the maxi¬ 
mum acceleration. 


Calculation of Springs—The maximum safe pressure of 
coiled steel wire springs and the deflection under a given 
load are obtained by means of the following formulae: 


rrr S d 3 

w=w Td-- 

c- 8 n P D '‘ 
F— -— 

Ed 4 


(90 A) 
(90 B) 


Where 

D = mean diameter of the coil, 

IV = maximum safe load in pounds, 

F- compression of spring, 
d = diameter of wire, 
n- number of coils in spring, 

5* = maximum safe fibre stress of material, 

E = torsional modulus of elasticity. 

P = load in pounds. 

The torsional modulus of elasticity may be taken at 12,000,- 
000 and the maximum safe working stress may be assumed 
to be 50,000 pounds per square inch. In general the outside 
diameter of the spring is made about seven-eighths the clear 
diameter of the valve, and the length of the spring under its 
initial pressure twice its outside diameter. These proportions 
can be used for guidance in laying down values for some of 
the factors in the above equations. 

To show the application of the above formulae we will cal¬ 
culate a spring for the 1 % inch valve above discussed. We 
will assume that the mean diameter (from centre to centre 
of wire) of the coil is i l / 2 inches. Since for a tangential cam 
the spring force must be somewhat greater than calculated 






240 


VALVES AND VALVE GEARS. 


for the uniform acceleration cam, we will calculate our spring 
so that it exerts 60 pounds pressure when the valve is closed 
and 80 pounds when the valve is lifted. The lift of the valve 
is Yz inch, and 20 pounds pressure therefore compresses the 
spring Yz inch. 

The necessary diameter of wire we find from equation 
(90A). Inserting the assumed values in this equation we have 

80 — 3 -H X 50,000 X d z 
8X1.5 

d = A/ —=0.183 inch. 
f 157,000 

This is approximately the diameter of a No. 7 (Birmingham 
gauge) steel wire, which we will adopt. The actual diameter 
of this wire is 0.18 inch. Now, to find the number of turns 
required, we take into consideration the fact that the total 
■deflection is to be Yz inch for 20 pounds load. 

Inserting in equation (90B)— 

y _ 8X 20 X i- 5 3 X n 
12,000,000 Xo. l8 4 

_ Yz X 12,000,000 X O. 18 4 

n 8 x 20 x 1.5 3 — = 9 a ^ roximaUl y • 

It was stated above that the length of the spring under its 
initial pressure is usually about twice its outside diameter. 
We will assume that the drawing in our case shows that the 
spring can be 3 inches long. Then, since the spring when 
the valve is closed is under 60 pounds pressure and it com¬ 
presses Yz inch for 20 pounds pressure, it follows that its 
initial compression is 

60 X /z __ inches. 

20 

The spring should therefore be wound to a length ot 

3 + lYz =4% inches. 

If the results obtained under the first assumptions are not 
satisfactory, different assumptions can be made and the calcu¬ 
lation made over again. In this connection it should be re¬ 
membered that for a given maximum pressure the stress to 
which the spring material is subjected varies directly with 
the diameter of the coil and inversely as the cube of the wire 
diameter. On the other hand, for a certain deflection per 
pound of load the number of coils, n, varies as the fourth 
power of the wire diameter and inversely as the cube of the 








VALVES AND VALVE GEARS. 


241 


coil diameter. Combining these two propositions, increasing 
the coil diameter increases the stress and decreases the num¬ 
ber of coils required, while increasing the wire diameter de¬ 
creases the stress and increases the number of coils required. 

In order to facilitate the design of coiled springs for poppet 
valves the maximum safe loads and deflections per coil have 
been calculated for all the wire sizes commonly used for 
valve springs and for the different coil diameters correspond¬ 
ing to each size, and the results are given in a table contained 
in the appendix. 

Spring Rests—A few foreign manufacturers bend one-half 
of the lowermost coil of the spring inward and pass it through a 
hole through the valve stem, as- shown at A, Fig. 143, but the 
general practice is to provide a cup-snaped washer for the spring 
to rest on. The construction shown at B is in very common use. 



Fig. 143.—Valve Spring Supports. 

The valve stem is provided with a transverse hole or slit in which 
is inserted a key. The key forms a support for the washer and is 
held from endwise motion by a downwardly turned flange on 
same. An upward flange on the washer centres the spring. In 
design C the valve stem is reduced in diameter near its lower 
end and a steel collar with a radial slot is slipped over this re¬ 
duced portion to support the washer, which is of the shape shown, 
and is stamped from heavy sheet metal. At D is shown an ar¬ 
rangement used to a considerable extent in France. It is identi¬ 
cal with B except for the fact that the sheet metal washer has 
no flange for retaining the key, which latter is somewhat wider 
and provided with 'a slot at the middle of its under side, whereby 
it is retained in position. 


















































242 


VALVES AND VALVE GEARS. 


Design of Push Rods—If the valves are located in side 
pockets of the cylinders, they are operated from the cams directly 
through the intermediary of a push rod. The latter usually con¬ 
sists of a cylindrical steel part moving in a brass guide secured 
to the crank casing. The push rod carries the cam roller or other 
type of cam follower at its lower end, and is provided with ad¬ 
justing means at its upper end. A typical push rod is shown in 
Fig. 144. 

In the design of these members, lightness is an important 
consideration, since the strength of the valve spring required, and 
the shock and noise produced, will be directly proportional to 
the weight of the parts moved by the cam, which include the 



Fig. 144.—Push Rod. 


push rod. In Fig. 144, A is the body portion of the push rod* 
which is drilled out from the top end for the sake of lightness.. 
The lower end is slotted transversely to receive the roller B,. 
which is carried on the roller pin C; the latter extends a short 
distance beyond the circumference of the push rod, its flattened 
ends working in narrow vertical slots in the push rod guide D, 
whereby the roller is kept parallel with the cam. At its upper 
end the push rod is internally threaded to receive the adjustingf 
plug E. The latter is provided with an hexagonal head, to which; 
a wrench can be applied, and a check nut F to lock it in adjust- 

































































VALVES AND VALVE GEARS. 


243 


ment. A sheet metal dust cap G is clamped between the lock nut 
F and the top end of the push rod body, and passes over the 
push rod guide, thus excluding dust and grit from the bearing 
surface. The plug E is also shown drilled out, with the object 
of reducing its weight. The projection at the top of the plug 
should be hardened, so it will not be marred or upset by the con¬ 
tinual hammering it receives as it strikes the end of the valve 
stem. 

The push rod guide is usually secured to the crankcase by 
means of two studs passing through lugs cast integral with it, 
and suitable nuts, or the guides of adjacent 
cylinders are held down by means of X- 
shaped yokes clamped to the crank case by 
means of a stud midway between the 
guides. 

A somewhat different design of push rod 
and guide is shown in Fig. 145. Here the 
bottom end of the push rod is made in 
the shape of a squared fork, and the bot¬ 
tom end of the guide is milled with a trans¬ 
verse slot, in which the square end of the 
push rod is guided. No special provisions 
are necessary for keeping the cam roller 
square with the cam. The adjusting means 
in this case consists of a plain capscrew 
and locknut. No provisions are made for a 
dust cap, and the top end of the guide is 
therefore left unfinished on the outside. 

Quite a number of designers provide, in 
addition to the valve spring, an extra spring 
on the push rod inside the guide, which is 
intended to always keep the cam roller in 
contact with the cam, so that the clearance 
will be found between the push rod and 
valve stem, instead of between the cam and 
roller. This has always seemed wrong to 
the writer. So far as the shock of lifting 
and closing is concerned, it does not matter much whether the 
clearance is at one point or the other, but if the shock takes 
place inside the crankcase it will certainly be less audible. An 
improved type of push rod designed to automatically take up the 
clearance is shown in Fig. 146. This push rod really consists 
of two parts, one telescoping the other. The upper one is in the 
form of a rod A, with an integral hexagonal flange near its mid- 



Fig. 145. 









































244 


VALVES AND VALVE GEARS. 


die, threaded at its upper end to receive the adjusting cap B and 
check nut, and cut down in diameter for a certain length near its 
lower end. This rod is a snug fit in a drill hole in the upper 
end of the lower portion C. A small coiled spring surrounds 
the rod A where it emerges from part C, and lengthens the push 
rod so as to take up the entire distance between the cam and the 
bottom end of the valve stem. Some oil is introduced into the 
bottom of the drill hole C, which acts as a cushion, and thus 
tends to reduce the shock. 


The need for figuring on a certain clearance 
in the valve operating mechanism arises from 
the fact that in the operation of the motor 
the cylinder is raised to a higher temperature 
than the valve stem and push rod, and hence 
expands more. Consequently, if the push rod 
is adjusted while the engine is cold so that 
there is practically no clearance in the valve 
operating mechanism, the parts will be found 
to clear a certain amount after the engine has 
been heated up. This difference in expansion 
is particularly pronounced in air cooled mo¬ 
tors, and is, no doubt, responsible for the 
fact that air cooled motors, as a rule, are 
somewhat more noisy than water cooled mo¬ 
tors. 

Cam and Roller Dimensions—The width 
of the cam can be made one-quarter the clear 
diameter of the valve when to be used with 
a roller follower, and one-third the clear di¬ 
ameter when to be used with a sliding type 
of follower. 

The cam roller should be kept fairly small 
in diameter, because with most designs of 
push rod the diameter of this rod must be 
slightly greater than the roller diameter. A 
roller which will revolve twice as fast as 
the cam will give satisfactory results, and the pin can be made 
of one-third the diameter of the roller. With mushroom type 
cam followers it is a good plan to offset the follower lengthwise 
from the middle of the cam, so the follower will rotate in its 
bearing and no groove will wear in its foot. Such a follower 
is shown in Fig. 147. 




Cam Levers—In order to reduce the side thrust on the push 


































VALVES AND VALVE GEARS. 


245 


rod, one armed levers are sometimes interposed between the cam 
and the push rod, as illustrated in Fib. 148. These levers are so 
arranged that they occupy a position perpendicular to the push 
rod axis when the valve is half lifted. The horizontal com¬ 
ponent of the cam thrust then comes on the lever and is taken 
up by the lever pivot. However, the push rod is not entirely 
relieved of side thrust, because the free end of_the lever moves 
in an arc or a circle instead of straight up and down, and the 
bottom end of the push rod has to slide on it. If that part 
of the lever on which the push rod bears is given such a curva¬ 
ture that the push rod is always at right angles to the contact 
surface, then the side thrust is equal to the product -of the 




Fig. 147. —Mushroom Type of Cam 
Follower and Guide Secured by Yoke. 


Fig. 148 —Lever Between 
Cam and Push Rod. 


upward thrust on the push rod by the friction coefficient. This 
is considerably less than the side thrust of the cam, but since 
the object of the cam lever is only partly accomplished and since 
it adds to the complication, this arrangement has lost considerable 
of its one time popularity. The aim now seems to be to provide 
the push rod guide with a liberal surface, bringing it down as 
near to the cam as possible and provide for copious lubrication. 

Overhead Valves—When the valves are located in the cyl¬ 
inder head with their stems pointing upwardly, they are operated 
by means of tappet rods (or tubes) and short tappet levers. A 








































































246 


VALVES AND VALVE GEARS. 

construction of this type is il¬ 
lustrated in Fig. 149. With 
these valve gears the weight of 
the moving parts is naturally 
greater than with ordinary 
valve gears for valves located 
in side packets, and every ef¬ 
fort should therefore be made 
to eliminate all unnecessary 
weight. The tappet rods are 
preferably made of steel tubing 
of an outside diameter of about 
y 2 inch. Into the lower end 
of the tube can be fitted an 
adjustable threaded steel rod 
whose lower end fits into a 
hole drilled into the push rod, 
and into the upper end a forked 
connector. The tappet lever 
on top of the cylinder is piv¬ 
oted on a bracket bolted to the 
cylinder head. A spring sur¬ 
rounding the lower part of 
the tappet takes up the slack 
in the mechanism outside the 

_ _ crank case, and adjustment for 

I'ig. 149-Overhead Valve Gear. wear> ^ can be ma<Je by 

screwing the rod at the lower end of the tube fur¬ 
ther in or out and locking it with a check nut. The 
forked connection between the tappet rod and lever should 
be provided with a liberal bearing with an oil hole, and the 
bracket bearing should also be liberal and provided with an oil 
cup. Sometimes the adjusting means in this form of valve gear 
consists of a set screw passing through the end of the tappet 
lever over the valve stem, which bears against the end of the 
valve stem and is secured by a lock nut. In the latest designs 
of this type the valves and tappet levers are covered by a 
housing. 

Integral and Built=up Camshafts—The great majority of 
the camshafts used in modern automobile motors are forged with 
their cams integral in the drop press and are turned and ground 
to size in special machines. Some makers, however, prefer to 
make the cams separately and to fit them to the shaft. This 
method came somewhat into disrepute in earlier years because 




































VALVES AND VALVE GEARS. 


247 


of inaccurate fitting and securing the cams merely by taper pins, 
so that they would occasionally come loose. At present, 
when the cams are made separate they are generally fitted by 
means of keys and held from endwise motion either by spacers 
on the shaft or by taper pins. The integral construction is, of 
course, a better manufacturing method, sipce it does away with 
fitting, but the built up construction has the advantage that it 
permits of hardening the cams more evenly and avoids danger of 
distortion. This method is in use by some of the leading 
manufacturers. 


Dimensions of Camshaft—The camshaft must be compara¬ 
tively stiff and must be well supported, so that it will spring or 
give very little under the shock it receives when opening the 
valves. This shock is, of course, greatest in the case of the 
exhaust valve, which must be 


lifted from its seat against 
the pressure of the gases in 
the cylinder near the end of 
the expansion stroke (about 
50 pounds per square inch, 
gauge). To this must be 

added the initial spring pres- 

♦ 

sure and the inertia of the 


I- 


i 


xl 





1 

-r— 



ZZZZ2 




i 


kzzZZZ 



Fig. 150. 


valve and push rod. In the case of the inlet valve only the 
last two items are to be counted with, as the valve does not have 
to be lifted against a pressure. The forte required to lift the 
exhaust valve may be assumed to be 75 pounds per quare inch of 
valve area. In Fig. 150 is shown a section of a camshaft of 
diameter d, whose length between the centres of adjacent bear¬ 
ing is /. Let the distance of the centre of the cam from the 
farthest bearing be x l. Then the flexure of the shaft at the 
centre of the cam under a total load P will be 

,(2* 4 +2Z 2 — 4* 8 ) 


y = 


PI* 


6 El 


(See Merriman, “A Textbook on the Mechanics of Materials,” 
Fourth Edition, page 77 ). The coefficient of elasticity may be 
taken at 30,000,000, and the moment of inertia I of a full circular 

T . Inserting these values in the above equation 


section is 


64 


we have 


PI' 1 


(2 x* + 2 x 2 — 4 x3 ) 


(91) 


* 8,800,000 d i 

This deflection y should not exceed 0.002 inch. For con 




































248 


VALVES AND VALVE GEARS. 


venience in using equation (91) the values of the expression in 
parentheses (=a) for values of x from to 1 inclusive are 
given in the following table: 


X = 0.5 

0-55 

0.6 

0.65 

0.7 

0 = 0.125 

0.122 

0.115 

0.104 

0.088 

x-0.75 

0.8 

0.85 

0.9 

0-95 

a= 0.069 

0.051 

0.033 

0.016 

0.005 


Suppose that in our 4x5 inch motor we provide three bearings 
for the camshaft, and that the distance between centres of bear¬ 
ings is 10 inches. If all of the cams are on one side, there will 
be four cams between each pair of supports. Luckily the ex¬ 
haust cams, on which comes by far the greatest thrust, will be 
close to the bearings, and it is a good plan to place them right 
alongside these. Supposing the bearing to be 3 inches long and 
the cam inch wide, the distance from the centre of the cam 
to the centre of the farthest bearing would be 8 % inches and the 
value of a would consequently be 0.04. The area of the valve 
port for a iX inch diameter is 2.4 square inches, and hence the 
valve thrust is 


2 -4 x 75 = 180 pounds 
Substituting in equation (91) we have 

180 X io 3 


0.002 


8,800,000 d 4 


X 0.04, 


hence 


d 


-vi 


180 X 0.04 X IO 3 O • 7 

=0.80 inch . 


.002 X 8,800,000 

However, since the inlet cams are located much farther from the 
supports, they may cause the shaft to flex more than the exhaust 
cams, even though the thrust on them is much less. The maxi¬ 
mum thrust on these cams may be taken as twice the mean 
spring pressure, say, 100 pounds. These inlet cams would be 
about 6 % inches from the farthest bearing, consequently 0=0.11. 
Inserting in the above equation 




IOO X °* II X !0 3 


= 0.89 inch . 


.002 X 8,800,000 

Hence the camshaft should be, say, ys inch in diameter, and its 
size is determined by the inlet cam thrust, rather than the 
exhaust cam thrust. 

Camshaft Bearings—Camshafts for four cylinder motors 
may be supported in either three, four or five bearings. Theo¬ 
retically it is possible to carry them in two bearings only, but 
this would require a very heavy shaft and offers no practical 
advantages. With T-head motors one bearing may be placed at 
either end of the camshaft and one between the first and second 







VALVES AND VALVE GEARS. 


249= 


and the third and fourth cams, respectively. In motors with 
individual or twin cylinders a bearing is sometimes provided on 
either side of the cams for each cylinder, making five bearings 
in all, but the most usual practice is to employ either three or 
four bearings. 

In earlier years it was customary to place the camshaft bearing 
between the crank casing and a bolted-on housing. At present, 
however, a sort of tunnel is generally provided for the cam¬ 
shaft (two in case of a motor with two camshafts) in the crank 
case casting, with annular supports for the camshaft bearings. 
When a built-up camshaft is used all of the bearings can be 
made in a single piece and slipped over the shaft as it is being 
assembled, but in case the camshaft is of integral construction 
the intermediate bearing or bearings have to be made in halves 
and bolted together. The bearings are made of an outer diameter 
slightly greater than twice the largest radius of the cam, so that 
when a shaft is assembled with its bearings the whole assembly 
can be pushed into place in the crank case from one end. The 
bearings are held in place by the point of a set screw passing 
through the wall of the crank chamber and entering a hole in 
the circumference. Most designs of camshaft bearings are pro¬ 
vided with oil pockets on top to catch the oil spray in the crank 
case, which arrangement is shown in Fig. 151. If there are no oil 
holes through the bearings, it is, of course, not essential that 
they should occupy any particular angular position, and they may 
then be held against lengthwise displacement by a set screw enter¬ 
ing a groove on their circumference, as shown in Fig. 152. The 
end bearings are sometimes flanged and bolted to the end wall of 
the crank case, which is provided with a boss for the purpose. 

The length of the camshaft bearings depends more or less upon 
the arrangement of the camshaft. Where three bearings are used 
on a four cylinder camshaft, the length of each bearing can be 
made about three times the diameter; where five bearings are 
used, each can be made from two to two and a half times the 
diameter. These bearings, for the sake of weight reduction, are 
made in the form of an inner and outer hub connected by a thin 
web. Occasionally ball bearings are used on the camshaft, but 
this practice is rare, owing to the fact that the camshaft runs 
at comparatively low speed, its load is light and its lubrication is 
effected automatically. Where such bearings are used they are 
mounted in the crank case in the same way as plain bearings. 

Manufacture of Cams— Camshafts with integral cams are 
drop forged from low carbon steel, and after being rough turned 
are cemented or carbonized at the cam portions, and then 





































































































































































VALVES AND VALVE GEARS. 


251 


quenched, so as to render the cam surfaces very hard and 
durable against wear. The cams are then ground in a cam 
grinding machine. 

Cams designed to be secured to their shafts are made either 
singly or in pairs (with a hub extending between the two cams). 
Such cams are generally made from tool steel and hardened all 
through. Separate cams are practically always used when only 
a comparatively small number of engines of a given design are 
to be built, say not over ioo, since in that case it does not pay to 
make the rather expensive dies for integral camshaft forgings. 
Of late, however, machines have been developed for either turn¬ 
ing or grinding integral camshafts from solid rolled stock, at 
nominal cost. 

A grinding machine aitachment for grinding individual cams 
is illustrated in Fig. 153, and it may be pointed out that cam 
grinding machines designed to handle integral camshafts work 
on practically the same principle. The cam to be ground is secured 
on an arbor inserted into a spindle which carries the master cam 
at its other end. The spindle is carried in a head provided with 
an upwardly extending arm whereby it is swung from the body 
or frame of the attachment, which is clamped to the grinder 
frame. The swinging head is forced to one side by a coiled 
spring so that the master cam presses against a stationary shoe 
plate. The spindle carrying the cams is revolved by means of a 
work driving pin on the face plate of the machine engaging with 
the forked end of a driving arm on it. In order to make the 
necessary allowance for wear in the grinding wheel, a set of 
different master cam shoe plates are furnished with the attach¬ 
ment, which are inserted one after another as the grinding 
wheel wears down, all of the plates having contact surfaces of 
different curvature. The master cam can also be cut by means 
of this attachment from a template. A method has recently been 
developed of grinding the cam with the flat sides of the grinding 
wheel, whereby inaccuracies due to change in curvature of the 
grinding surface with wear are eliminated. 

For grinding an integral camshaft it is necessary to first make 
a master camshaft with all of the cams correctly spaced angularly 
and longitudinally. This camshaft is driven in synchronism with 
the camshaft to be ground, and acts on a roller carried by a cross 
slide carrying the grinding wheel, the roller being drawn against 
the master cam by a spring acting on the cross slide. 

The cam rollers and pins are generally made from tool steel 
and hardened. 

I11 motors for pleasure vehicles great importance is now at- 


252 


VALVES AND VALVE GEARS. 





Fig. 153. —Landis Cam Grinding Attachment. 































































































































VALVES AND VALVE GEARS. 


253 


tached to silent operation of the valves. The factors which 
enhance silent valve action are as follows: Small valve lift, 
light weight of lifted parts, careful design of cam outline so as 
to lift and seat the valve gradually, accurate adjustment of 
push rod for clearance and a stiff and rigidly supported cam¬ 
shaft. Valve housings filled with grease further enhance silence. 

Compression Relief Cams—With large motors it is de¬ 
sirable to reduce the compression when starting the motor by 
means of the crank, since it takes a great deal of physical ef¬ 
fort to turn over a large, multi-cylinder, high compression 
motor by hand against full compression. On some motors 
all of the cylinder pet cocks are connected together and can 
be opened simultaneously from a position convenient to the 
operator when about to crank the motor, whereby the effort 
required in cranking is, of course, greatly reduced. However, 
the more approved plan of relieving the compression for 
starting consists in providing the exhaust cams with an extra 
cam protuberance at one end and nearly opposite the main 
cam protuberance, which holds the exhaust valve open dur¬ 
ing the first part of the compression stroke. 

The exhaust camshaft is so arranged that it can be slid 
lengthwise in its bearings, and the cams are made consider¬ 
ably wider than the width of the cam roller, so that when 
the camshaft is in its normal position the cam roller does 
not touch the relief cam. The camshaft is connected to 
an operating lever from which connection is made to 
a handle or other operating device in front of the 
car. A spring normally holds the camshaft in such a posi¬ 
tion that the relief cams are out of line with the cam rollers, 
but when the shaft is pulled forward by the operator the re¬ 
lief cams come in line with the cam rollers and the exhaust 
valves are held open not only during the regular exhaust 
period, but also during the first part of the compression 
period. The sides of the relief cam protuberances which the 
cam roller must mount are suitably inclined. A portion of 
an exhaust camshaft with relief cams and operating mechan¬ 
ism is shown in Fig. 154. At the forward end of the cam¬ 
shaft there is a short shaft, which is so connected to the 
camshaft that it does not need to turn with it, but that the 
camshaft must follow it whenever it is moved longitudinally. 
This short shaft is surrounded by a spring, which tends to 
keep the camshaft in its normal position. The forward end 
of the short shaft projecting through the cam gear housing 
cover is flattened and pivoted to an operating lever. The cam- 


254 


VALVES AND VALVE GEARS. 


shaft should be fitted with a ball thrust bearing to take up the 
end thrust due to the spring. 

Another common arrangement consists in surrounding the 
rear end of the camshaft with a coiled spring, which bears 
against the rearmost cam and the rear wall of the crank case 
respectively. One of the cams is formed with a groove for 
a sliding collar, with which engages a shipper lever secured 
to an operating shaft extending through the side wall of the 
crank case. 



Fig. 154. —Compression Relief Mechanism. 


Concentric Valves.—Fig. 155 shows a design of concentri¬ 
cally arranged inlet and exhaust valves as used by the H. H. 
Franklin Manufacturing Co. in its air cooled motors for a num¬ 
ber of years. At present no such valves are being used in auto¬ 
mobile motors, so far as the writer is aware, but they are coming 
into extensive use on aeronautic motors. One advantage of these 
valves is that they permit of large valve opening areas in con¬ 
nection with a combustion chamber of cylindrical or nearly hemi¬ 
spherical form, which is desirable from the standpoint of thermal 
efficiency. Besides, the exhaust valve is cooled by the fresh 
charge coming in through the inlet valve, and is, therefore, not 
so easily pitted by the hot gases passing through it. The Frank¬ 
lin company discarded these valves apparently because they 
proved noisier than ordinary valves. 

Auxiliary Exhaust Ports.—To insure the quickest possible 
escape of the exhaust gases, the cylinders are sometimes provided 



















































VALVES AND VALVE GEARS. 


255 


with ports, known as auxiliary exhaust ports, which are uncovered 
by the piston just before it reaches the bottom end of the stroke. 
These ports are opened very quickly and allow the greater portion 
of the spent gases to escape before the piston completes its down 
stroke. Before it was discovered that the exhaust valve had to 
begin to open some 40 to 45 degrees before dead centre in order 
to insure the prompt clearing of the cylinder, and when one or 
two cylinder motors were the rule, these auxiliary exhaust ports 
were very extensively used, since it was found that if the exhaust 
valve opened only a little before the end of the stroke these ports 
considerably increased the engine power. However, when four 
cylinder engines came into use, a difficulty cropped up. When 




Fig. 155.— Franklin Concen- Fig. 156— Franklin Auxiliary 
tric Valves. Exhaust Valve. 

the burning gases are discharged through the auxiliary exhaust 
port of one cylinder into an exhaust pipe common to all cyl¬ 
inders, they may flow back through the auxiliary exhaust port 
of one of the other cylinders whose piston is at the moment at 
the end of its inlet stroke and has therefore also uncovered its 
auxiliary exhaust port. In that case the charge in the latter 
cylinder may be fired, straining and heating the engine without 
doing an} r useful work. In order to prevent this it is necessary 
to place a valve outside each of the auxiliary exhaust ports which 
will shut off communication between the port and the common 
exhaust pipe at the end of the inlet stroke. 








































































256- 


VALVES AND VALVE GEARS. 


At present the auxiliary exhaust port is used only on aircooled 
motors, where it has the great advantage that, since it permits 
the hot gases to escape quickly, it reduces the heating of the 
cylinder walls and particularly of the exhaust valve. The H. H. 
Franklin Manufacturing Co., which has used these auxiliary 
valves on its air cooled motors (see Fig. 156), states that only 
about 30 per cent, of the total quantity of exhaust gas passes 
through the regular exhaust valves, which begin to open only at 
the beginning of the exhaust stroke. These valves can, there¬ 
fore, be made smaller than would otherwise be permissible; 
they need not be lifted against any pressure, and the strain on 
them and the consequent noise are much reduced. These ad¬ 
vantages are obtained, however, at the cost of some complication, 
since an extra valve per cylinder is required. But these valves 
for the auxiliary ports need not be absolutely tight and need not 
be lifted against the pressure in the cylinder; consequently they 
do not require much care. The auxiliary ports should be made 
comparatively narrow, so they will not reduce the effective length 
of the power stroke too much, and be made to extend nearly 
half way around the cylinder, being divided by so-called bridges 
into a number of separate ports so the piston rings will not catch 
on the edges of the port. In racing and aeronautic motors the 
auxiliary ports often are in the form of holes drilled through the 
cylinder wall, the spent gases being discharged directly into the 
atmosphere. 

Single Cam Valve Gear—It is possible to operate both the 
inlet and the exhaust valve of each cylinder by means of a single 
cam. This construction, which was first described in The Horse¬ 
less Age of February 5, 1902, gained considerable prominence 
through its use on Fiat racers in 1906. The Fiat design is illus¬ 
trated in Fig. 157. It will be seen that the two valves are located 
in the cylinder head at an angle of 30 degrees with the cylinder 
axis and are held down to their seats by a multiple leaf spring. 
They are opened by means of a rocking lever pivoted to a bracket 
screwed into the centre of the cylinder head, one end of which 
lever connects to a valve rod extending down the side of the cyl¬ 
inder and connecting with the cam follower at its lower end. This 
valve rod is surrounded by a coiled spring which is stronger than 
the flat spring holding the valves to their seats. It will be noticed 
that the cam has both a cam protuberance for opening the exhaust 
valve and a flat which allows the valve rod to descend below its 
.normal position and the coiled spring to open the inlet valve 
against the pressure of the leaf spring. A similar valve arrange 


257 


VALVES AND VALVE GEARS. 

ment was used on Pope-Toledo stock cars in this country for 
some years. It has been urged against it that it does not permit 
of overlapping of the inlet and exhaust opening periods, but this 
overlapping does not seem to be desired anyhow. 

F 1 "- 157 also illustrates another idea in valve construction which 



Fig. 157. —Fiat Single Cam Mechanism With Ported Inlet 

Valve. 

has been resorted to in racing motors in order to increase the 
power output. This is the ported inlet valve. It will be observed 
that the inlet valve head is provided with ports, so that the charge 
passing through the valve seat, instead of all passing around the 
outside of the valve head, comes partly around the circumference 































258 


VALVES AND VALVE GEARS. 


and partly through the ports in the valve head. This design is a 
product of speed competition and has been used on racers only. 
Of late it has become customary to provide two inlet and two 
exhaust valves per cylinder in big racing motors. 

Enclosed Valves—Lately it has become customary to enclose 
the entire valve mechanism so as to protect it against dust and 
grit, and muffle its noise. En bloc motors frequently have their 
cylinders cast with a housing containing the valve springs and 
push rods, which can be closed by means of a cover held in 



place by studs screwed into bosses cast on the cylinders, and 
nuts. Such a valve casing protects all of the more delicate parts 
of the motor, and gives the motor that “ironclad” appearance 
which has long been a feature of electric motors of the street 
railway and automobile types. A typical design of such a valve 












































































































VALVES AND VALVE GEARS. 


259 


housing is illustrated in Fig. 158. This motor is of the en 
bloc type, with integral inlet and exhaust passages. It has a 
flange cast along the under side of the valve box, and another 
flange extends sideways from the lower end of the casting. The 
latter flange, together with the valve box on top and two side 
walls at the ends of the cylinder block, forms a housing for the 
valves, which is closed by a large ribbed cover plate cast of 
aluminum. 

Such valve housings are also occasionally provided in motors 
with twin cast cylinders, but the arrangement is then less ad¬ 
vantageous, since four side walls are required, and the total 
weight of the housing is naturally greater. In order to attain 
practically the same end with less weight, protecting sleeves 
are sometimes placed over the valve springs. A design of such 
sleeves is shown in Fig. 159. Each valve spring is surrounded 
by two sleeves, the upper smaller diameter one screwing into 
the lower one. By partly unscrewing the smaller sleeve from 
the larger one, the ends of the sleeve are forced against flanges 
on the valve stem guide and the push rod guide respectively, 
whereby a dust-tight joint is insured. In putting the sleeves in 
place the smaller one is placed inside the larger one, and the 
two are held against the valve box from below, when the valve 
may be inserted and the spring and its retaining washer put in 
place, after which the larger sleeve may be lowered to its regu¬ 
lar oosition. 


CHAPTER XII. 


CAMSHAFT AND ACCESSORIES DRIVES. 

The camshaft receives its motion from the crankshaft, from 
which all of the accessories must also be driven. These acces¬ 
sories generally include a magneto, a water pump, an oil pump 
and a fan, and sometimes also a timer, an air pump and a light¬ 
ing dynamo. The camshaft and accessories drives are generally 
combined, all of the parts deriving their motion from a single 
pinion on the crankshaft, so it will be advantageous to discuss 
the whole subject under one head. 

There has for some time been a gradual evolution in the de¬ 
sign of camshaft and accessories drives, which is still in 
progress. The object aimed at is quiet operation of the drives, 
combined with the most compact arrangement, and the greatest 
accessibility of the accessories. Until recently it was the gen¬ 
eral custom, where camshafts ran parallel with the crankshaft, 
to drive them by spur gears, and to have small spur gears for 
the magneto and water pump mesh with the camshaft gears. 
About two years ago helical camshaft gears were widely intro¬ 
duced on account of their more silent operation, and now the 
silent chain bids fair to come into use for this purpose for the 
same reason. In Europe it is now customary to drive the mag¬ 
neto and water pump by means of a cross shaft in front of the 
motor, driven from the camshaft through a pair of helical gears, 
and the same practice is finding favor in this country. 

Arrangement of Drive—In Fig. 160 are shown a number 
of different arrangements of the drive for one or two camshafts 
located in the crank chamber, and for the chief accessories. A 
depicts an arrangement which is now practically obsolete, but 
was used a good deal in the early years of four cylinder motors. 
Each of the two camshafts is provided with a large diameter 
spur gear, meshing directly with a pinion on the crankshaft. 
(It may be pointed out that in this, the same as in all other dia¬ 
grams in Fig. 160, the crankshaft is indicated by a large black 


260 



261 


CAMSHAFT AND ACCESSORIES DRIVES. 

circle, and the camshafts are indicated by small black circles.) 
With the right hand camshaft gear meshes a small gear for 
driving the pump and magneto. On a four cylinder motor the 
magneto must be driven at crankshaft speed, and the magneto 
gear therefore has the same number of teeth as the crankshaft 
pinion. The reason that this arrangement of camshaft gears has 
been given up is that it necessitates the use of gears of very 
large pitch diameter, resulting in very high pitch line velocities, 
and it is very difficult to make gears running at a high pitch line 
velocity operate silently. 

For T-head motors an arrangement like that shown at B, or 
one similar to it, is now generally used. The pinion on the 
crankshaft meshes with an idler gear of equal or slightly greater 
pitch diameter, which in turn meshes with the two camshaft 
gears. A comparison with arrangement A shows at once that 
the gears and pinions of arrangement B are much smaller, and 
with an equal engine speed the pitch line velocity is much less. 
In this case the magneto is located on the left hand side of the 
motor, and the pump on the right hand side, and each has to be 
driven by a separate gear meshing with one of the camshaft 
gears. 

At C is shown a camshaft drive by means of silent chains as 
successfully used by the Humber firm in England on T-head 
motors. There are two driving sprockets on the crankshaft, 
which connect by a silent chain each with a driven sprocket 
on each of the two camshafts. The right hand camshaft carries 
in addition to the sprocket through which it receives its motion 
a driving sprocket for transmitting motion by a third silent chain 
to the driving shaft for the pump and magneto. At D is shown 
a silent chain drive for a motor with all valves on one side, and 
with the pump and magneto located on the side opposite to the 
valves. 

Diagram E shows an arrangement of the camshaft and ac¬ 
cessories drives which has lately come into considerable favor 
abroad. Most of the new European motors are of small size, 
and of the L-head type. The camshaft is driven from the crank¬ 
shaft directly through a pair of spur gears. On the camshaft is 
mounted a helical gear meshing with another helical gear on a 
cross shaft above it, which drives the magneto at one end and 
the pump at the other. 

At F is shown the camshaft drive employed on the Winton 
Six. The crankshaft pinion meshes directly with the camshaft 
gear, which latter in turn meshes with a small gear for the 
magneto and pump, and with an idler transmitting motion to a 


262 


CAMSHAFT AND ACCESSORIES DRIVES. 


driving gear for the fan. This being a six cylinder motor, the 
magneto is required to run at times crankshaft speed, and 
the magneto gear, therefore, has only two-thirds the number of 
teeth of the crankshaft pinion. The fan is driven at twice 
crankshaft speed. 



Fig. 160.—Diagrams of Camshaft and Accessories Drives on 
Engines With One or Two Camshafts in Crank Case. 

In L-type engines both the magneto and pump are generally 
located either at the valve side, or at the ends of a cross shaft 
in front, but in a few instances they are located on the side op¬ 
posite the valves, and a drive like that shown at G, with an 












CAMSHAFT AND ACCESSORIES DRIVES 263 

idler between the crankshaft pinion and the magneto and pump 
gear may then be used. 

Engines with overhead camshafts are very rare in this country, 
but are used to quite an extent in Europe, in Germany particu¬ 
larly on truck motors. Two arrangements of the camshaft drive 
for motors with overhead camshafts are shown in Fig. 161. 



That shown at A is employed on a German truck motor. There 
is a vertical shaft at the front of the motor driven from the 
crankshaft by a pair of mitre gears, and driving the camshaft 
through a set of bevel gears in the ratio of i to 2. A cross¬ 
shaft for the magneto and pump is driven from the vertical shaft 
through a pair of helical gears at a ratio of 1 to 1. The vertical 
shaft also drives the fan through a pair of bevel friction wheels, 
and in addition carries the governor which, however, is not 
shown in the sketch. 

At B is shown the silent chain drive for the German over¬ 
head camshaft engine. One silent chain runs from a sprocket 


















264 


CAMSHAFT AND ACCESSORIES DRIVES. 


on the crankshaft to a sprocket on a short shaft on the side of 
the cylinders from which the water pump and magneto are 
driven, and a second silent chain runs from a second sprocket on 
this short shaft to the sprocket on the camshaft. The first three 
sprockets are of equal size, but that on the camshaft has twice 
the number of teeth as the others. Both of the overhead cam¬ 
shaft driving gears here described are completely enclosed. 

Dimensions of Camshaft Gears —Since the camshaft or 
shafts must rotate at one-half the angular speed of the crank¬ 
shaft, the driven gear must have twice the number of teeth as 
the crank shaft pinion. If spur or helical gears are used, the 
teeth are generally of eight pitch, and the gears are made of a 
width of face 



— inch 
8 


(92) 


This may be increased by one-eighth to one-quarter inch in the 
case of truck or other motors intended for services in which 
weight is not of very great importance. The driving pinions 
are made of steel, and the camshaft gears of cast iron or 
bronze. Unless they are carefully made, these gears may be a 
source of considerable noise. It is with the object of reducing 
this noise that gears cut with helical teeth having a tooth angle 
of 30 degrees or thereabout have been largely substituted in 
recent years for plain spur gears. The form of the web of the 
gears also has some effect on the noise. Occasionally the cam¬ 
shaft gears are cast with regular spokes, but more generally 
they are formed with a web of about one-quarter the thickness 
of the face of the gear, and with a number of large circular holes 
evenly spaced. 




Fig. 162. B. 


The most secure 
method of fasten¬ 
ing the camshaft 
gear to the shaft 
consists in provid¬ 
ing the shaft with 
an integral flange 
and bolting the 
gear to it, and this 
method is em¬ 
ployed in a num¬ 
ber of motors used 
on cars of the 
highest grade (see 
Fig. 162, A). How¬ 
ever, the more 








































CAMSHAFT AND ACCESSORIES DRIVES. 265 

common method consists in fitting the gear by means of 
one or two Woodruff keys against a shoulder on the 
shaft, either on a cylindrical or a tapered seat. When a 
tapered seat is used the shaft is always drawn to it by means of 
a special nut of comparatively small thickness. Such a nut is 
sometimes also used in connection with a cylindrical seat, as 
shown at B, Fig. 162, though it is not necessary if the gear is 
prevented from drifting on its shaft by means of the housing. 
Again, the gear may be fastened to its shaft by means of one 
or two taper pins. It is essential, of course, that the gears be 
fitted to the shaft with great accuracy, so far as their angular 
relation is concerned, so that the valves may be properly timed. 
Some manufacturers provide an adjustment by making the cam¬ 
shaft gear with a separate flanged hub, and bolting the web of 
the gear to the flange of the hub by three bolts, the bolt holes 
in one of the parts being oblong so as to permit of angular ad¬ 
justment. 

Silent Chain Camshaft Dri ves—When silent chains are 
employed for driving camshafts, the inch pitch size is usually 
selected, though for small engines or for driving accessories 
only, the still smaller pitches made by one or two of the manu¬ 
facturers of these transmission devices will serve. As regards 
the width of the chains to be used, the manufacturers of the 
Renold chain recommend a one-half inch pitch chain 1.4 inches 
wide for engines of 4 inch bore or less with a single camshaft, 
and 1.2 inches wide for double camshaft engines. For engines 
of more than 4 inches bore chains of the same pitch, and of 
1.4 and 1.7 inches width for single and double camshafts re¬ 
spectively, are recommended. If a smaller size is selected, pro¬ 
vision must be made for adjustment, on account of the more 
rapid wear of the smaller chains. In the design of silent chain 
drives for camshafts it is desirable to provide for chains with 
an even number of links, so as to avoid bent or cranked links at 
the joint. 

A simple method of accurately timing the valve gear (which 
can also be used where the camshaft drive is by spur or helical 
gears) is illustrated in Fig. 163. The sprocket on the camshaft is 
provided with a dummy hub, and the flange on this hub and the 
web of the sprocket are each provided with three oblong holes 
inclined against each other, through which are passed clamping 
bolts. By removing these clamping bolts farther in toward or 
out from the camshaft the sprocket will be moved angularly with 
relation to the camshaft, and thus the timing changed. 


2056 CAMSHAFT AND ACCESSORIES DRIVES. 

Magneto and Pump Couplings—Both the magneto and the 
water pump have their own shafts, which must be connected to 
the driving shaft or shafts by some form of coupling. In the case 
of the magneto it is essential that the drive be absolutely syn¬ 
chronous, that is, without angular variation, which would dis¬ 
turb the timing of the ignition. In the case of the pump, on the 
other hand, a slightly flexible coupling is considered advan¬ 
tageous by some designers, for the reason that if any solid 
particles should accidentally get into the pump, a flexible drive 
would not be injured as easily, and if the pump should ever 
freeze up the flexible member might protect it against injury 
when the motor is turned over. In the case of both drives, it is 
well that the coupling allow of a slight disalignment of the two 



shafts connected, since it is very difficult to get these shafts ab¬ 
solutely in line with each other. 

In Fig. 164 A shows a magneto coupling consisting essentially 
of a yoke with slotted ends fastened to the driving shaft, and 
a hub fastened to the magneto shaft with opposite radially ex¬ 
tending arms engaging into the slots of the yoke on the driving 
shaft. The yoke is secured to the driving shaft by means of two 
setscrews, which allow the magneto to be accurately timed. A 
similar design suitable for a pump drive is shown at B in the 
same figure. Here one of the members is replaced by a set of 
saw blade steel springs which are clamped to the slotted end of 
one of the shafts, and engage with the slotted ends of the yoke. 

The form of coupling most extensively used for this purpose is 




















































CAMSHAFT AND ACCESSORIES DRIVES. 267 



Fig. 164. —Pump and Magneto Shaft Couplings. 


the Hookham joint, shown at C. This consists essentially of 
three discs, the two outer ones of which are formed with hubs 
and keyed or pinned to the driving and driven shafts, respec¬ 
tively. The adjacent faces of these discs are cut with diametrical 
slots, and between the two discs is located a floating disc with 
tongues on opposite sides at right angles to each other, which 
tongues exactly fit into the grooves in the outer discs. It will 
readily be seen that this form of coupling permits the two shafts 
to be out of centre without causing any binding in the bearings. 





Fig. 165. —Adjustable Flanged Coupling. 


























































































268 


CAMSHAFT AND ACCESSORIES DRIVES. 

A simple form of magneto coupling is that shown at D, which 
consists merely of two flanged hubs secured to the driving and 
driven shafts, respectively. One of the flanges is cut with a 
radial slot, and the other is provided with a substantial pin en¬ 
gaging with the slot. The slot is of the exact width of the 
pin, and of such depth that the pin does not bottom in it. 

Some designers merely use very long driving shafts of com¬ 
paratively small diameter, which 
they couple to the magneto or pump 
by means of ordinary flanged coup¬ 
lings. A flanged coupling permitting 
of very close angular adjustment of 
the magneto is shown in Fig. 165. 
The two flanges are drilled with 
holes closely spaced all around their 
circumference, one of them having 
two holes more than the other. The 
flanges are secured together by means 
of two bolts passed through holes at 
opposite ends of a diameter, and by 
moving these bolts from hole to hole 
the angular relation of the driving 
and driven parts will be slightly 
changed. Thus if one flange has 20 
and the other 18 holes, the driven 
member can be advanced or retarded 
in steps of 2 degrees relative to the 
driving member. 

Oil Pump and Timer Drive— 

The oil pump is generally located 
at the side of the oil well at the 
bottom of the crank chamber, and 
is most conveniently driven through 
a vertical shaft. The timer, on 
the other hand, should preferably 
be located high up at the side of 
the cylinder, so that it will be ac- 
Fig. 166. —Oil Pump and cessible and the connections from 
Timer Drive. it to the spark plugs will be short. 

It is, therefore, also driven by a vertical shaft, and in quite 
a number of cases the same shaft serves for the oil pump and 
the timer. Fig. 166 illustrates the arrangement. The timer 
must be driven at camshaft speed, and hence the gear ratio of 
the pair of helical gears by means- of which it receives its mo- 
































































CAMSHAFT AND ACCESSORIES DRIVES. 269 

tion from the camshaft must be 1:1. The driving pinion must 
be of small pitch diameter because it has to pass through the 
camshaft tunnel. This makes it difficult to get the centre of the 
vertical shaft sufficiently far from the centre of the camshaft, so 
that all parts will clear the crankcase, and in order to obviate 
this difficulty some designers use two sets of helical gears for 
the vertical shaft drive, employing a short horizontal inter¬ 
mediate shaft. The timer is also occasionally driven by bevel 
gears from the rear end of one of the camshafts. Some form of 
universal coupling, usually a simplified Hookham joint, is inter¬ 
posed between the drive shaft and the pump shaft. The timer, 
as a rule, has no shaft of its own, but is made with a hub which 
fits over the driving shaft at its top end. It must not be over¬ 
looked that when the camshaft carries helical or bevel pinions 
it is subject to end thrust and should be provided with thrust 
bearings, preferably of the ball type. 

Fan Drive—The great majority of cars are equipped with 
a fan back of the radiator, which must be driven from some 
moving part of the engine In a few cases the fan is driven by 
spur or bevel gears, in which case a friction clutch must be in¬ 
serted in the drive in order to prevent straining of the trans¬ 
mission when the motor is speeded up or slowed down abruptly, 
because the average fan has considerable inertia. The gear drive 
can, of course, be entirely enclosed and run in oil, but for ordi¬ 
nary purposes it is rather too complicated. 

The usual method of driving the fan consists in the use of a 
flat belt connecting from a pulley on the fan to a larger pulley 
on either the crankshaft, camshaft or magneto shaft. The 
favorite location for the fan driving pulley is on the magneto or 
pump shaft, which runs at twice the speed of the camshaft, and 
hence requires a pulley of only one-half the diameter as one 
mounted on the camshaft. Of course, if the pulley was mounted 
on the crankshaft it would require to be no larger, but belts do 
not operate as advantageously in a vertical position as in a 
horizontal or inclined position. The fans usually run at from one 
and one-half to twice the engine speed, the latter being the more 
usual. The pulleys can be designed to give a belt speed of 1,800 
feet per minute at normal engine speed. The width of the belt 
should depend, of course, upon the size and form of the fan, 
but for a rough approximation it may be taken as 

tv — A 4 - 1 inch. 

8 

Trapezoidal belts are also used for driving the fan, running 
on V groove pulleys. One advantage of these belts is that, since 


270 CAMSHAFT AND ACCESSORIES DRIVES. 


they are wedged in between the wide walls of the pulley, they 
do not require to be as tight as a flat belt in order to transmit 
the same amount of power, hence they are not so likely to stretch 
and slip. On the other hand, the wedging action absorbs a certain 
amount of power and the bending of the comparatively thick 
belt is also attended with losses, so, as far as efficiency of trans¬ 
mission is concerned, it is hardly likely that the trapezoidal belt is 
equal to the flat belt. 


CHAPTER XIII. 


CRANK CASE AND OILING SYSTEM. 

The crank case in an automobile motor serves as the main 
structural part which carries the cylinders and the crankshaft 
and is in turn supported upon the vehicle frame. It also forms 
a housing for most of the important working parts of the en¬ 
gine, protecting them against dust and splashing mud, and it 
performs important functions in connection with the lubrication 
of the motor. 

The crank case generally forms a substantially cylindrical hous¬ 
ing sufficiently large in diameter to enable the crankshaft with 
the connecting rod heads to rotate freely within it and sufficiently 
long to accommodate all of the cylinders of the engine. It is 
provided with circular openings over which the cylinder castings 
are bolted, with bearings for the crankshaft and camshaft, and 
with supporting arms by means of which it is carried on the 
main frame or a subframe. 

Materials—Practically all crank cases for pleasure car motors 
are cast of aluminum, which is used on account of its low 
specific gravity. In reality, not pure aluminum, but alloys con¬ 
taining a large percentage of the metal are used, as pure aluminum 
has not sufficient strength for the purpose. One of these alloys, 
containing about 8 per cent, of copper, has a tensile strength of 
18,000 pounds per square inch. Another, containing 3 per cent, 
of copper and 15 per cent, of zinc, has a tensile strength of 22,000 
pounds per square inch. The latter (specific gravity = 3) is, how¬ 
ever, somewhat heavier than the former, requires more care in 
casting and is more liable to shrinkage strains, which in part 
makes up for its somewhat greater tensile strength. One or two 
makers use manganese bronze for that part of the case which 
carries the crankshaft bearings and has the supporting arms cast 
integral with it. This alloy has a tensile strength of from 65,000 
to 70,000 pounds per square inch, an elastic limit of about half 
these figures, and a specific gravity of 8.5 The crank cases of 


271 



272 


CRANK CASE AND OILING SYSTEM. 


commercial vehicle motors are generally made from cast iron, 
because the reduction in weight which it would be possible to 
secure by means of aluminum does not compensate in this case 
for the additional cost. In motors of the double cylinder opposed 
type the crank cases are sometimes parted in the middle and each 
half is cast integral with its respective cylinder, and in four 
cylinder en bloc motors the upper half of the crank case is also 
sometimes cast integral with the cylinder block, in which cases 
these parts are necessarily of iron. 

One and Two Cylinder Crank Cases—Crank cases for single 
cylinder vertical motors are generally made in halves joined in a 
plane perpendicular to the crankshaft axis. The crank cases 
of double cylinder opposed motors can be arranged in different 
ways. It has already been pointed out that occasionally they 
are parted in a vertical plane and each half is cast integral with 
the corresponding cylinder. They may also be cast in halves 
parted in a vertical plane through the crankshaft centre, with 
the cylinders bolted on. Then, they may be cast in halves parted 
in a horizontal plane or cast in box form with one or two remov¬ 
able circular end plates. 

A crank case design for a double cylinder opposed motor is 
shown in Fig. 167. This type of motor is used chiefly for light 
commercial vehicles, and the crank case is made of cast iron. The 
main portion of the case is cast in a single piece, with openings on 
opposite sides for the cylinders—an opening at one end through 
which the crankshaft can be introduced, and an opening on top 
through which the camshaft with its gear may be put in place. 
In double cylinder opposed motors the valves are generally 
placed on top. It was formerly customary to carry the crank¬ 
shaft and push rods on a removable top plate, so the whole valve 
mechanism could be removed together, but this construction is 
patented, and since the patent on it has been sustained it has 
been given up by some manufacturers. Now the crank case is 
generally provided with square slots on top into which the cam 
shaft bearings fit and which are closed by the top plate. The 
push rod guides are forced into bosses cast integral with the case. 

When one end of the crank case is cast integral the opening 
at the other end must be of a diameter slightly greater than the 
largest diameter of the crankshaft. With such a large opening 
the flange of the end plate comes uncomfortably near the bear¬ 
ing boss for the camshaft, unless the camshaft is placed at a 
greater distance from the crankshaft than would otherwise be 
necessary. To avoid this difficulty, removable end plates can be 
used on both ends and the crankshaft “wriggled” into place be- 


CRANK CASE AND OILING SYSTEM. 


273 


fore either end plate is bolted to the case, in which case the end 
plates can be of smaller diameter. Another problem that arises 
with engines of this type, especially if of relatively long stroke, 
is that the distance between the crankshaft and camshaft is 
greater than the required distance between the cylinder centre 
and valve centre. This is due to the fact that in this type of 
engine the line connecting the crankshaft and camshaft is at right 
angles to the cylinder axis, whereas in a vertical motor it makes 
angles of from 45 to 60 degrees with that axis. This makes it 
necessary to either place the valve chamber farther from the 
cylinder axis, thereby increasing the length of the valve passage, 



and, consequently, the compression space wall area, or to offset 
the valves from the camshaft centre and provide the push rod 
with a sort of overhanging striker. 

Four and Six Cylinder Crank Cases—In crank cases for 
four or six cylinder vertical motors there has been a gradual 
evolution from the earliest crude designs, which made it prac¬ 
tically impossible to adjust the crankshaft bearings and difficult 
to get at the crank pin and wrist pin bearings. Fig. 168 shows a 
design which was once much used, but is now obsolete. The 
crankshaft bearings are formed between the lower and upper 
halves of the case, and the upper half is cast with the supporting 
arms and with hand holes through which it is possible to get at 
the connecting rod cap bolts for the purpose of adjusting them. 








































































274 


CRANK CASE AND OILING SYSTEM. 


This design of case is, of course, very rigid, but, on the other 
hand, its manufacture involves a number of difficulties. The 
boring tool cannot be observed by the workman while the in¬ 
termediate bearing is being bored, that is, just when the strain 
on the boring bar is greatest. In fitting and scraping in the 
bushings the case must be repeatedly taken apart and assembled 
again, which adds to the cost of manufacture. The only way 



Fig. 168. —Bearings Between Halves of Case. 


in which wear in the crankshaft bushings can be taken up is by 
placing shims between the bushing and the bearing hub—a rather 
unsatisfactory method. In order to get at the crank case bear¬ 
ings, the crankshaft, flywheel, connecting rod and pistons have 
to be dropped, and altogether the adjustment of the bearings of 
this type of case is a most difficult problem. 

A somewhat similar design, in which the crankshaft bearings 
are formed between the two halves of the case and the sup¬ 
porting arms are cast on the lower half, is used by such well 
known firms as the German Daimler Company and the American 
Locomotive Company. The general design is illustrated in Fig. 
169. This type of case possesses special advantages where the 
motor is mounted on the car in such a way that the bearings 
cannot be gotten at from underneath, as where it is located over 
the front axle. In order to take up and scrape any of the crank¬ 
shaft bearings, the upper half of the case and cylinders can be 
removed together as a unit. The workman then has access to all 
of the principal working parts and can work upon them with 
comfort. The pistons, piston rings and connecting rods can be 
inspected and changed without disturbing the crankshaft bear¬ 
ings. In adjusting and scraping the crank pin bushings the con¬ 
necting rods can be swung through a half circle. If the crank- 



























CRANK CASE AND OILING SYSTEM. 


275 



shaft bearings require adjustment the lower half of the case need 
not be disturbed, unless a new bushing is needed. 

What is undoubtedly the most extensively used type of crank 
case at the present time is illustrated in Fig. 170. All of the 
crank bearings are entirely supported from the top half, and the 
lower half of the case serves only for the protection of the 
working parts against dust, etc., and to contain a supply of 
lubricant. This makes it possible to run the motor for short 
periods with the lower half of the case removed, so all of the 
bearings can be thoroughly examined and any necessary adjust¬ 
ments made with the greatest convenience. When this form of 
construction was first adopted the lower half was made in the 
form of a sem-cylindrical pan with partition walls between the 

























































276 CRANK CASE AND OILING SYSTEM. 

crank pits from i to 2 inches high, so that not all the oil would 
run to one end of the case when the car was ascending or 
descending a grade, but lubricating systems have recently under¬ 
gone considerable change, and what is known as the pump cir¬ 
culating system is now almost universally used. This involves 
the use of oil troughs directly underneath each crank, in which 
the oil is kept at a constant level by means of overflow pipes, 
and an oil well or sump beneath these troughs in which the over¬ 
flowing oil accumulates. The oil trough and the oil well may 
both be cast integral with the lower half of the crank case, but 
the more usual plan is to form the troughs in the lower half 
and make the oil well a separate casting bolted to the lower 
half. Another alternative consists in casting the oil well integral 
with the lower half and inserting inrto it the troughs, which in 
that case may be made of sheet metal. In small motors the 
entire oil well is sometimes made of pressed steel. 



In a considerable number of engines the crank cases are made 
in barrel form ; that is, the mam portion of the case is made a 
single casting, with openings for circular bearing plates, for the 
cylinders and for inspection purposes. A case of this type is 
shown in Fig. 171. The advantage of this construction is that 
with a certain weight of metal it gives the most rigid case. The 
inspection holes may be either at the side or in the bottom, but 
if there is an oil well at the bottom they must be at the sides. 
The only valid objection which can be raised against this form 
of construction is that in order to remove the crankshaft it is 
necessary to take the engine out of the car. Some crank cases 
of this type have been made of such relatively large size, and 
with such large handholes on the side, that piston and connect¬ 
ing rods could be removed through these handholes without dis- 
































































CRANK CASE AND OILING SYSTEM. 


277 

turbing other parts of the motor, but in general it would be 
considered that the end did not justify the means. 

General Design—The length of the crank case depends, of 
course, upon the size and form of the cylinders, crankshaft and 
crankshaft bearings. The size of the crankshaft and its bear¬ 
ings can be determined by means of the rules laid down in the 
chapter on Crankshafts. The crank case must allow of the free 
rotation of the crankshaft with the connecting rod heads, which 
calls for a substantially cylindrical form. However, if the cam¬ 
shafts are to be placed in tunnels formed in the casing it is 
necessary to depart somewhat from a cylindrical form, since 
these shafts must necessarily be outside the circle described by 
the outermost points of the crank, and if the crank case were 
to be made a cylinder of such a diameter as to enclose the cam¬ 
shaft, it would be unduly bulky. Moreover, with fairly long 
connecting rods it is advantageous to bring the flange to which 
the cylinder bolts considerably higher above the crank case centre 
than the maximum radius of the crankshaft. The oil well at 
the bottom also alters the external contour of the case, but need 
not affect the cylindrical form of the crank case proper. 

The cylindrical form is advantageous not only because of its 
weight economy, but on account of its strength. The crank case 
does not lend itself well to mathematical treatment, but must be 
designed by the use of good engineering judgment. The main 
stresses to which it is subjected are as follows: When the explo¬ 
sion takes place in one of the cylinders the cylinder, owing to 
the reaction of the explosion on its head, tries to pull away from 
the crank case, exerting a stress on the cylinder bolts, the crank 
case bosses into which they are secured and those parts of the 
crank case between these bosses and the lower halves of the 
crankshaft bearings which take the direct thrust of the explosion. 
The torque reaction of the motor attacks the crank case also 
through the cylinder bolts and strains particularly the supporting 
arms and the parts intermediate between the cylinder flanges and 
the arms. Any “weaving” of the frame tends to break off the 
arms and twist the crank case. In the design of the crank case 
we must carefully consider each of these different stresses. 

In the first place the cylinders must be firmly secured to the 
crank case. It is not well to depend upon screw threads in 
aluminum for this connection, and one or the other of the 
fastening means shown in Fig. 172 is usually employed. The 
construction shown at A involves the use of a special stud bolt 
turned with a collar at its middle, which collar enters a counter¬ 
bore in the boss for the stud. The stud is bolted to the case by 


278 


CRANK CASE AND OILING SYSTEM. 


means of this collar and a castellated nut. Rules for the size 
of the stud were given in the chapter on The Cylinder. The 
lugs of the aluminum casting should have a thickness at least 
two and one-half times the diameter of the stud. Arrangement B, 
Fig. 172, employs a special double diameter bolt. This has a 
thread in the aluminum to hold it in place when the cylinder 
casting is put on the crank case, but the thread is not depended 
upon to take the strain of the explosion. The lugs for the 
outermost bolts generally come near the end walls of the case 
and a reinforcement or rib can be provided between them. 
Similarly, the lugs for the bolts at the middle of a four unit 
cylinder block will come at the same place as the central parti¬ 



tion wall. If there is no partition wall, as in the cases of four 

cylinder motors with two bearing crankshafts, arch-shaped stiffen¬ 
ing ribs can be run across the under side of the top wall of the 
crank case, as shown in Fig. 173. Such stiffening ribs are some¬ 
times also provided in other types of motors between cylinder 
openings where there is no partition wall. These stiffening ribs 
compensate for the weakening effect of the bulge in the case 
necessitated by the camshaft tunnel. 

In divided crank cases there is usually one partition wall for 
each intermediate crankshaft bearing. This partition wall is in 
that part which carries the bearings and in both halves if the 
bearings are “pinched” between them. Ordinarily these parti¬ 
tion walls are provided with vertical ribs on both sides and 
with a flange at the bottom (see Fig. 170). Occasionally, how 
ever, in order to secure still greater rigidity, they are made ot 
box girder form. That is, there are two partition walls a short 










































CRANK CASE AND OILING SYSTEM. 


279 



distance apart, which are stiffened by a vertical connecting web 
at their centre. As shown in Fig. 174, this form of girder makes 
a very substantial bearing support. 

Thickness of Sections—The thickness of the metal (alumi¬ 
num) around the cylinder openings should be % inch per inch 
of cylinder bore. The side walls of the case are generally made 
flr of an inch thick, which is about the smallest thickness that 
can be successfully cast where the pattern has such a large sur¬ 
face. Around the joint between the upper and lower halves a 
flange is run which substantially doubles the thickness of the 
metal there. This flange is made of the same height as its width. 
If all the bearings are supported by the upper half the two halves 
can be joined by means of t* inch bolts for engines up to 4V2 
inches bore, and % inch bolts for engines of larger bore, spaced 
4 to 5 inches. These bolts should pass through lugs twice as high 



1 _ n 


1 

- -j 



■L-' ■ 

— e- 

' | 

-e-— 


f-4- 


L 


j 




\ 


Fig. 174 .—Box Girder Construction. 
































































































280 CRANK CASE AND OILING SYSTEM. 

as the bolt diameter and of sufficient size to accommodate the 
bolt head and nut. Gaskets of tough brown paper soaked in oil 
are used between the cylinders and the crank case and between 
the separate parts of the crank case in order to make these 
joints oil-tight. 

Supporting Arms—Vertical motors, as a. rule, are cast with 
four supporting arms, two on each side, at the front and rear, 
respectively. Occasionally, however, they are provided with what 
is known as the three point support, in which case they have 
only two supporting arms and are in addition supported on a 
cross member of the frame at either the front or-rear crankshaft 
bearing. Integrally cast supporting arms should preferably be 



B 


A 


Fig. 175.— Methods of Fastening Engine Arms to Frame Bars. 

so located that one of their sides is a direct continuation of an 
end wall of the casing, whereby greater strength is insured. The 
arms are generally of channel section with the open side below; 
where they join the crank case, if the latter be of the divided 
type, they are generally of the full depth of that part of the case 
with which they are cast, and thence they taper down to the 
height of the frame section or less, if they are secured to 
the main frame. Motors are now generally supported directly 
upon the main frame, and the conventional method of securing 
their arms to the frame side bars is shown at A in Fig. 175. 
The channel section of the frame where the supporting arms 
come is filled with wood, the ledge on the supporting arm is 
placed on top of the frame and a bolt is passed through the 
frame section, the wood filling and the outer wall of the sup- 
oorting arm. When the motor is carried on a subframe (made 










































CRANK CASE AND OILING SYSTEM. 281 

either of angle or channel steel) the ledges on the arms are made 
larger and provided with one or more bosses for bolts passing 
\ertically through them and through the subframe member. An 
arm end intended to be supported on a subframe is shown at B 
in Fig. 175. The metal in the arms is generally made about 34 
inch thick, and y 2 inch where the arm bolts to the frame. 

Separate Supporting Arms —The long supporting arms make 
the crank case casting somewhat difficult to handle in machining, 
for which reason some manufacturers prefer to use separate 
supporting arms of cast steel, manganese bronze or drop forged 
steel. Separate arms are quite popular with manufacturers who 



make engines for the trade, since they have to build their engines 
to fit frames of different widths, which could not easily be done 
if the arms were cast integral. Separate supporting arms are 
generally either of I or channel cross section and are secured 
to the forward and rearward end of the crank case, respectively, 
by means of two large bolts at each end, passing through lugs on 
the end wall of the crank case, sometimes on either side of the 
outer crankshaft bearing hubs. Some foreign manufacturers 
support their engines by means of steel tubes which are passed 
through suitable lugs cast on the end walls of the crank case 
and provided at their ends with flanged hubs which are bolted to 
the frame side bars. 

Breathers —Four cylinder vertical motors, as well as several 
other types, have to be provided with so-called breathers through 
wm'ch air may escape when the volume of the crank chamber is 
reduced owing to the unequal speed of the pistons during the 






















282 


CRANK CASE AND OILING SYSTEM. 


upper and lower halves of their stroke, respectively. If no 
breathers were provided an air pressure would be created in the 
crank chamber which would tend to force the oil contained in 
the chamber through the bearings and the joints of the crank 
case. The breathers, of which there are generally two, are fre¬ 
quently placed on the supporting arms, which in that case are 
so formed that the crank case chamber extends partly into them 
(see Fig. 176). The reason for placing the breathers on the 
arms is that they should be connected to a part of the crank 
case in which there is comparatively little oil spray, so that as 
little oil as possible will be forced out with the air. Another 
reason is that these breathers generally serve the purpose of oil 
filling tubes, and when they are placed on the arms they are 
somewhat farther removed from the cylinders and are, therefore, 
more accessible from the outside. 



In the design of the breathers provision has to be made for 
separating the oil contained in the air and causing it to flow 
back into the crank chamber. A widely used form of breather 
is shown at A, Fig. 177. This is in the form of a cast tube with 
downwardly inclined separator partitions extending into the tube 
alternately from opposite sides. The oil globules striking the 
under side of these walls will adhere to them and when suffi¬ 
cient oil has accumulated it will run back into the case. The 
dome shaped cover further tends to collect the drops of oil and 
t.o return them to the case, while the air escapes through the 




























































CRANK CASE AND OILING SYSTEM. 


283 


narrow space between the tube and cover. At B is shown a 
design of breather tube with somewhat different means for sepa¬ 
rating the oil globules from the air. At the bottom of the 
breather tube proper is inserted a sort of thimble with perforated 
wall. Near the top of the tube there is a wire gauze strainer 
which tends to keep back the oil globules, and the outlet from the 
tube is through numerous small holes in the side wall of the cap. 
At C is shown an arrangement used by several French designers. 
This consists simply of a leather flap valve at the top of the 
tube which permits any excess of pressure in the crank case to 
equalize itself. This is based on the idea that the excess pres¬ 
sure in the crank case is due to leakage of gas past the pistons 
and rings, for which an outlet must be provided if the air and 
gas are not to blow the oil through the bearings. 



Cam Gear Housing—The camshaft and accessories driving 
gears are in nearly all motors located at the forward end, but 
occasionally they are placed at the rear end. These gears must 
be completely enclosed in order that they may be effectively lubri¬ 
cated and protected against dust. In a barrel type of crankcase 
the housing can be made in two parts, one formed integral with 
the crank case or its end plate, and the other one bolted to it. 
In crank cases of the divided type one part of the gear housing 
is cast with the lower half of the case and another with the upper. 
That part with the lower half can be in the form of a semi- 
cylindrical chamber of the proper size to accommodate the crank¬ 
shaft pinion, the pinion being introduced with the crankshaft from 
above, or it may be left open in front. The parts of the housing 
are flanged at their edges and bolted together so as to form an 
oiltight joint. Some designers make practically the whole hous¬ 
ing integral with the crank case and make the detachable part in 



















































284 


CRANK CASE AND OILING SYSTEM. 


the form of a cover plate, which is sometimes of pressed steel, 
while others divide the housing substantially in a central plane. 
In a few instances the gear housing is made separate from the 
crank case, divided in a plane passing through the crankshaft and 
camshaft axes, and clamped to the forward crankshaft and cam¬ 
shaft bearing hubs. The parts of the housing are held together 
by about ten cap screws, for which lugs must be provided in the 
castings. These lugs may be cast on the outside walls, but it is 
preferable to cast them on the inside walls, since this makes the 
outside surface smoother and easier to keep clean. A typical 
design of cam gear housing is shown in Fig. 178. 

Starting Crank Support—At the present time nearly all 
pleasure cars are provided with self-starters and no starting 
crank is carried in position at the front of the car as formerly. 
However, a starting crank is still carried in the tool box for 
emergencies and a bearing must be provided for it at the forward 
end of the motor. This is either formed integral with or bolted 
to the cam gear housing, or it is fastened to the forward cross 
member of the frame. It is now usual to make the engaging 
ratchet of the starting crank of the same diameter as the shaft 
itself, so the crank can be inserted into its bearing from in front. 
The front of the bearing is ordinarily closed by means of a cap 
to prevent dirt and water working into it. 

However, the “permanent” starting crank has not been entirely 
discarded, being still used on practically all motor trucks. If the 
shaft is of the same diameter as the ratchet, it may be provided 
with a circular groove, into which engages the point of a 
spring pressed plunger or a spring pressed steel ball carried 
in a pocket formed on the bearing, to prevent the crank from 
jarring out. Generally two adjacent grooves are cut into the 
shaft of the starting crank, so located that the starting crank 
ratchet is in engagement when the plunger or ball is in one 
groove and just out of engagement when it is in the other groove 
(see Fig. 178.) The spring referred to enables the crank to find 
its proper place when it is thrown off by the starting motor. A 
somewhat different arrangement is also widely used, in which 
the starting crank is pressed forward against a thrust bearing 
by means of a coiled spring surrounding it. When it is desired 
to engage the starting crank the operator must force it back 
against the pressure of this spring, which is comparatively light, 
however. 

Fan Bracket—In many motors the fan bracket is secured 
to the cam gear housing, which is provided with one or two 
horizontal flats or bosses for the purpose, to which the bracket 
is bolted, or that portion of the housing cast integral with the 
crank case may be formed with an upwardly extending ribbed 
bracket, to which the fan bracket is bolted. Fig. 179 shows a 
typical construction. When this type of fan bracket support is 


CRANK CASE AND OILING SYSTEM. 285 

used the belt tension is usually adjusted by means of an eccen¬ 
tric bushing in the fan bearing hub or around the fan stud, as 
shown in the illustration. 

Bearing Caps—As already stated, in divided crank cases the 
bearings are generally formed between one-half of the case and 
c^ps bolted to same. If the caps form the lower half of the 
bearing they take the direct thrust of the explosion, while if they 
form the upper half they are subject to the force of inertia of the 
reciprocating parts only, but since at racing speeds this is sub¬ 
stantially the same as the force of the explosion, the caps and 
their retaining studs and bolts must be the same in either case. 
The caps are held to the integral parts of the bearing hubs by 



Fig. 179.—Fan Bracket 
Bolted to Cam Gear 
Housing. 




Fig. 180.—Crank Bearing Cap 
and Locking Means. 


either two or four bolts or studs, four being used if the bearings 
are relatively long, on motors of larger size or of the better 
grade. The rear bearing cap in a motor with all bearings sup¬ 
ported from the top half of the case, which has to carry the 
weight of the flywheel, should be provided with four bolts or 
studs. The studs of the caps should be made of such size that 
their combined cross-sectional area is equal to one-twelfth the 
projected surface of the bearing. As in the case of the cylinder 
studs, it is not well to depend upon threads in the aluminum, 
and the studs should therefore be backed up by nuts. A method 
for securing the nuts of the cap studs in place is illustrated in 





























































































286 


CRANK CASE AND OILING SYSTEM. 


Fig. 180. A stamping is made of sheet brass, curved at the 
middle to the shape of the cap, with two hexagonal holes near 
its ends fitting snugly over the nuts, and a small round hole 
at its centre through which passes a pin cast integral with the 
cap, and which takes a split pin that holds the stamping in place. 
Some designers provide the bearing caps with ribs to help carry 
off the heat generated by friction in the bearing. 

The integral part of the bearing and its cap are separated by 
shims when first assembled, so that when the bearing wears it 
can be readjusted by removing one or more of the shims from 
each side of each bearing. When the bearing hub is made in 
halves the bushing is also made that way and with flanges on 
both ends preventing endwise dislocation. Angular dislocation 
is prevented by one or two pins through the cap and bushing, 
or if the bushing is of the die cast type a pin may be cast in¬ 
tegral with it, which enters a hole into the cap from the inside. 

The bearing metal generally used is some white brass or 
babbitt composition, the bushing being either cast of this com¬ 
position or cast of bronze and lined with the composition. The 
bearing bushings are drilled with oil holes and oil grooves in the 
usual way. 

The camshaft bushings are cast of bearing bronze, and if 
the camshaft does not extend beyond the rear bearing the 

latter is generally closed 
at its outer end, so that 
no oil will be lost through 
it. It may either be cast 
“blind’’ at one end or a 
piece of sheet brass may 
be soldered over the end. 

Oil Guards —To pre¬ 
vent the loss of oil 
through the front and 
rear crankshaft bearings, 
the crankshaft is provid¬ 
ed with oil guards at 
these places, as already 
explained. These consist 
of collars integral with 
the shaft turned to 
a sharp edge from 
which the oil is 



Fig. i8i.—Oil Guard. 



































CRANK CASE AND OILING SYSTEM. 


28 ? 


flung off when the crankshaft turns at considerable speed 
An annular recess must be formed in the end of the bearing 
hub, as shown in Fig. 181, to catch the oil thus flung off and 
return it to the bottom of the crank chamber. If the end 
bearing is supported by the upper half of the crank chamber the 
lower half is formed with a sort of pocket equal in outside di¬ 
mensions to the bearing hub, into which the oil thrown off by the 
oil guard drops. If, on the other hand, half of the bearing 
is in the lower half of the crank case a connecting passage must 
be drilled from the oil guard recess to the crank chamber 

through a rib at the bottom of 
the bearing hubs. Some manu¬ 
facturers use felt washers or 
stuffing boxes at the outer 
ends of bearings to prevent 
the egress of oil and the in¬ 
gress of dirt and grit. 

Magneto and Pump 
Brackets—The magneto is 
generally set on a bracket cast 
integral with the upper half of 
the case, suitably strengthened 
by ribs. In designing this 
bracket it is well to remember 
that most foreign magnetos 
measure 45 millimeters from 
the base to the armature shaft 
centre and are provided with 
four >>£-inch dowel holes in the 
base located at the corners of 
a square with 50 millimeters 
sides; in the largest models, 
now rarely used, the height of 
the shaft is 50 millimeters and the dowels are at the corners 
of a rectangle measuring 50 by 80 millimeters. The magneto 
is generally secured in place by means of two flexible steel bands 
anchored to the bracket and provided with a bolted joint between 
them, which makes the magneto very easy to remove. This is 
illustrated in Fig. 182. The magnetos are, however, also pro¬ 
vided with four threaded holes in the base and can therefore 
also be secured to the bracket by means of cap screws. 

Most of the water circulating pumps are manufactured by the 
car manufacturers themselves, and their bases or supporting 
members are of many different shapes. Sometimes the same 








































288 


CRANK CASE AND OILING SYSTEM. 


kind of bracket is provided 
for the pump as for the mag¬ 
neto, while again the pump, if 
of the centrifugal type, may be 
provided with an extension of 
its radial flange by which it is 
bolted to a supporting bracket. 
Power tire pumps also are now 
often fitted to engines ana 
standardization of brackets and 
shaft height is at present under 
way. 

Inspection Hole Covers— 

Certain forms of crank cases 
have inspection holes in their 
sides and it is a good plan to provide their covers with 
some form of quick acting locking mechanism rather than 
to secure them in place by machine screws. One form of 
such locking mechanisms consists merely of a yoke swivel¬ 
ing on a stud secured into the crank case, whose ends en¬ 
gage between projections cast near the centres of adjacent 
cover plates. (Fig. 183). Another arrangement is shown in 
Fig. 184. This cover is of rectangular shape and is held in 
place by means of sliding bolts with tapered ends engaging 
into openings formed by bent pieces of steel strip riveted to the 
wall of the case on the inside. 



Fig. 183. —Circular Inspection 
Hole Cover 



Fig. 184. —Rectangular Inspection Hole Cover. 








































































CRANK CASE AND OILING SYSTEM. 


289 


On the whole very few automobile motors are now provided 
with inspection holes, there being generally little room on either 
side of the motor in a car. 

Continuous Web Construction —Some manufacturers, par¬ 
ticularly those turning out high-priced cars, cast continuous webs 
on the crank case, extending between the frame and the crank 
case and the whole length of the latter, or between the front and 
rear supporting arms. The advantage of these webs is that 
they prevent road dust from working up to the cylinders, valve 
springs and other motor parts, tending to keep the parts 
cleaner. Fig. 185 shows the continuous web construction on the 
Owen motor. 



Fig. 185.—Continuous Web Construction. 

Splash Lubrication —So many different systems of engine 
lubrication have been worked out that it is difficult to classify 
them. We may distinguish, however, between the following: 
Plain splash; splash from constant level troughs with pump 
circulation; part force feed and part splash; force feed without 
splash; gravity feed without splash. 

The simple splash system was much used on earlier models of 
cars. The crank case contained a supply of oil into which the 
connecting rod head dipped at each revolution of the crank, 
thereby splashing the oil over all the interior parts. As the oil 
worked out through the bearings or past the pistons the oil level 
in the crank chamber fell and the plash became weaker. When 
the operator considered that the oil level had become too low he 
would replenish the supply by either pouring oil into the crank 
chamber through a filling hole or transferring it from a supply 
tank to the crank chamber by means of a hand pump provided 
for the purpose. 










290 


CRANK CASE AND OILING SYSTEM. 


The disadvantages of this system are apparent. The oil level 
varied constantly and so, naturally, did the rate of supply of 
lubricant to the bearing surfaces. When the crank chamber was 
replenished with oil the supply to the cylinder walls was generally 
excessive and the motor in consequence would pour dense smoke 
out through the muffler. Some makers found it necessary, in 
order to reduce this smoking, to provide a baffle plate between 
the cylinder and crank chamber with an opening in the form of 
a rectangular slot for the connecting rod to pass through. More¬ 
over, the lubrication required frequent attention from the operator, 
and if the latter was careless the motor could easily be seriously 
injured through lack of oil. 

Mechanical Oilers—Through these deficiencies splash lubri¬ 
cation came into disrepute, and lubrication by means of multiple 
feed mechanical oilers came into vogue. These oilers consisted 
of prismatic oil reservoirs containing a set of plunger pumps 
forcing the oil through individual oil leads to all the different 
parts requiring lubrication. There was at least one pump foi 
each lead, and oil was forced mechanically to each bearing, in 
proportion to the speed of the engine. While this system was 
quite effective, its complication was objectionable. The oiler 
usually had a belt drive, which involved an element of uncer¬ 
tainty, and the maze of oil tubes (sometimes as many as twelve) 
running from the dashboard along the side of the engine, ren¬ 
dered access to the latter difficult and gave it a frail and non- 
mechanical appearance. Moreover, there was an adjustment for 
each feed and the chances of mal-adjustment, and, consequently, 
of improper lubrication of some bearing, naturally increased with 
the number of feeds. It was felt that a simpler system was 
needed and one that was at the same time thoroughly automatic. 
It was then that the pump circulating system was introduced. 

Gravity and Force Feed—It will be seen from the above 
that there are two methods of feeding oil to the bearing sur¬ 
faces, either by gravity or by means of mechanical pressure. The 
latter may be as high as io pounds per square inch, or even 
higher. When oil is forced to the bearings under considerable 
pressure there is evidently less danger of it being squeezed out 
by the pressure on the bearing. It is generally agreed that with 
pressure oil feed the specific pressure on the bearing can be 
made greater and the bearing will last longer than with gravity 
feed. On the other - hand, the pressure feed system is more 
expensive to install and also somewhat less economical of oil. 
This system is best adapted for motors which pre required to 


CRANK CASE AND OILING SYSTEM. 291 

work at or near their full load capacity for long periods at 
a time. 

The oil may also be fed to the bearing surfaces by gaseous 
pressure derived either from the engine exhaust or an air pump. 
This system enjoyed some popularity about the same time that 
mechanical lubricators were in vogue, but is not much used now. 

Force Feed Non=Splash System—In this system oil is drawn 
from a reservoir, generally located at the bottom of the crank 
case by means of a pump and is forced to the main bearings 
of the crankshaft. A branch from the oil delivery pipe leads to 
the dash, where it connects to an oil gauge indicating the pres¬ 
sure acting upon the oil. After leaving the pump the oil may 
either pass through a distributor which connects the pump suc¬ 
cessively with each of the main bearings, or it may divide be¬ 
tween leads to the individual bearings which are connected in 
parallel. The latter arrangement is probably not quite as de¬ 
pendable as that in which a distributor is used, for the reason 
that if one of the leads should become clogged up all the oil 
would pass through the remaining leads and practically no 
pressure would be exerted to dislodge the obstruction. A pres¬ 
sure regulator in the form of a by-pass valve is connected to the 
delivery side of the pump. 

In bearings carrying a uniform or nearly uniform load it is the 
rule to feed the oil to the side opposite to where the pressure 
comes. In the crankshaft, crank pin and piston 
pin bearings of a gasoline motor the pressure 
constantly changes from one side to the other, 
and it does not seem to matter whether oil 
is fed to the under or the upper side, both 
methods being successfuly employed. 

The oil passes to the main bearings through 
oil tubes and passages drilled in the crank 
chamber wall. The bearing bushing is also 
prvoided with a drill hole from which radiate 
the oil grooves. A radial hole in the crank¬ 
shaft registers with the hole in the bushing 
once every revolution of the crankshaft and 
allows oil to enter the drilled passage in the 
crankshaft. A groove may be cut in the bear¬ 
ing part way around the circumference to keep 
the oil passage in the crankshaft open for a 
longer time. In drilling the crankshaft, some 
manufacturers provide oil pasages only from Tube to Piston 
each of the main bearings to the adjacent Pin. 



Fig. 186.— Oil 











292 


CRANK CASE AND OILING SYSTEM. 



u^r 


[ 


i 


































































































































































































































































CRANK CASE AND OILING SYSTEM 


293 


crank pin bearings, while others drill the crankshaft from end to 
end, so that in case the oil lead to one of the main bearings 
should become obstructed oil would be fed to that bearing 
and the adjacent crank pin bearings through the other leads. 
The oil is carried outward to the crank pin bearings by centrif¬ 
ugal force, in addition to the pump pressure acting upon it. 
The piston pin bearing, which is one of the most difficult bear¬ 
ings to lubricate satisfactorily, is supplied with oil from the 
crank pin bearing through a hole drilled through the centre of the 
connecting rod or through a tube passing up alongside the con¬ 
necting rod or through a hollow connecting rod. (See Fig. 186.) 

In this system oil is fed at a comparatively fast rate and a good 
deal of oil works through the bearings and is thrown off the 
crank arms, etc., onto the interior cylinder, piston and crank 
chamber walls. This oil is depended upon to lubricate the cylin¬ 
der, the camshaft bearings, the interior members of the valve 
gear, etc. All of the oil thus thrown off, except that which 
reaches the flame swept surfaces of the cylinder wall and is 
burned, returns to the bottom of the crank case and collects in 
the oil reservoir, from which it is drawn by the pump. At the 
pump inlet there is a strainer which separates all metal dust 
or particles of dirt from the oil before it is recirculated. The 
Marmon oiling system, illustrated in Fig. 187, is of this type. 

The above described system is the extreme mechanical system 
in which the largest number of parts receive their lubricant di¬ 
rectly by mechanical pressure. From it can be evolved a num¬ 
ber of other systems by increasing the number of bearings oiled 
by the oil spray and correspondingly reducing those oiled by 
force feed. Thus, for instance, the piston pin bearings may be 
oiled by means of the spray by providing the small end of the 
connecting rod with oil pockets on top, as described in connec¬ 
tion with the connecting rod. Then, instead of feeding the oil 
from the crankshaft main bearings to the crank pin bearings 
through oil ducts all the way, the oil working through the main 
bearings at their inner ends may be caught in oil rings, as shown 
in Fig. 188, and from these conducted through diagonal holes to 
the crank pin bearing surfaces. Such a diagonal hole, by the 
way, is not as objectionable as a radial hole, since a radial hole 
considerably reduces the amount of stock at a certain section— 
and that usually at the point of maximum stress—thus tending 
to concentrate the strain and considerably weakening the shaft. 
In the case of crankshafts for engines of short stroke as com¬ 
pared with the bore it is posible to drill inclined holes through 
the crankshaft main bearing crank arm and crank pin, thus 


294 


CRANK CASE AND OILING SYSTEM. 

largely reducing this weakening effect and at the same time doing 
away with plugged holes, which are apt to develop leaks. 

Gravity Feed Non-Splash System—In the gravity feed non¬ 
splash system an oil reservoir is carried somewhere under the 
engine hood, generally near the cylinders, so that the oil will be 
kept in its normal state of fluidity in extremely cold weather by 
the heat radiated by the cylinders. The reservoir or tank may be 
placed slightly above the cylinder heads and oil is fed from it by 
gravity to each of the crankshaft bearings through oil tubes and 
holes drilled in the crank case wall. From the main bearings 
the oil flows through the hollow crankshaft to the crankpin 
bearings, and possibly through tubes to the piston pin bearings, 
the rest of the bearings being lubricated by the spray. The 



greater part of the spray drains back to the bottom of the crank 
case and thence to an oil well below it, from where it is returned 
to the oil tank by the pump. The oil is filtered at several points, 
as at the inlet to the pump and at the inlet to the feed pipes, and 
the oil tank is provided with an oil gauge to indicate the amount 
of oil on hand. The Pierce-Arrow lubricating system, illustrated 
in Fig. 189, is of this class. 

Combined Force Feed and Splash System—This is one of 
the most popular methods at the present time. The oil is carried 
in a sump or oil well at the bottom of the crank case from which 
it is drawn by the pump and forced through a circulation indicator 
on the dash to the crankshaft main bearings, and sometimes also 
to the cam gear compartment. The oil working through the 
main bearings drains to the bottom of the crank case, which is 

























CRANK CASE AND OILING SYSTEM. 295 

formed with troughs into which splashers on the connecting rod 
heads dip at each revolution of the crankshaft. The oil level in 
the troughs is kept constant by means of overflow standpipes 
or holes in the side wall of the case through which the excess oil 



drains off to the sump or oil well below. The Columbia lubri¬ 
cating system, illustrated in Fig. 190, is of this type. In this 
system the oil is forced by the pump into an oil passage formed 
in the top wall of the crank chamber and running the entire length 
thereof. From this main oil passage the oil is conducted through 































































































































































































296 


CRANK CASE AND OILING SYSTEM. 


short tubes to the crankshaft main bearings, the cam gear hous¬ 
ing and the central camshaft bearing. The tube leading to the 
rear main bearing leads through the indicator on the dash. It 
will be observed that all of the bearings are provided with splash 
pockets, so that lubrication would not cease even if the pump 
should fail to work. 

One of the difficulties with the splash systems in the past has 
been to prevent excessive lubrication of the cylinder walls and 
consequent smoking of the motor. That is, if the oil level is 
carried sufficiently high to insure that all of the bearings receive 



Fig. 190.—Constant Level Splash Circulating System 

(Columbia). 

an adequate supply of oil the cylinder walls will receive an ex¬ 
cessive amount, with the result that the motor will smoke. This 
it has been attempted to remedy in the past by means of baffle 
plates between the cylinders and crank case and by means of 
scraper rings at the bottom end of the piston which are designed 
to scrape off a portion of the oil thrown onto the cylinder walls 
The trouble has been greatly reduced since it has become cus- 










































































































































































































































CRANK CASE AND OILING SYSTEM. 297 

tomary to grind both cylinders and pistons, which gives a better 
fit and does not allow the oil to get by the piston so easily. 
Some manufacturers cut one or two oil grooves in the wall of 
the piston, with the upper side perpendicular to the cylinder axis. 
The upper edge of the groove scrapes the oil from the cylinder 
wall, and the oil collecting in the groove returns through holes 
drilled through the piston wall to the crank chamber. 

Pump Feed and Splash System—Some manufacturers, in¬ 
stead of carrying the supply of oil in an oil well, carry it in a 
tank placed near the cylinders from which oil is pumped into 
the separate compartments of the crank chamber at the rate it is. 
consumed by the motor. The rate of feed of the pump or pumps 
is adjustable. This, therefore, is a splash non-circulating system.. 



Fig. 191.—Constant Level Air Displacement System. 

(E-M-F.) 


Air Displacement Constant Level System—In another 
system oil is carried in a reservoir at the side of the crank case 
and cast integral therewith. From the bottom of this reservoir 
lead two downwardly inclined pipes into the crank chamber 
compartments, with their mouths at the height at which it is 
desired to carry the oil level in these compartments. The oil 
reservoir is air tight, and when the mouths of the two pipes are 
covered with oil (that is, when the oil in the crank chamber 
stands at the proper level) the reservoir is “air bound” and no 
oil can flow into the crank case. When, however, the oil level 
sinks below the mouths of the tubes, air bubbles will rise through 
the tubes to the top of the reservoir and an equivalent amount 
of oil will flow into the crank chamber, until the mouths of the 
















































298 


CRANK CASE AND OILING SYSTEM. 


tubes are again covered with oil. The amount of oil in the 
reservoir is indicated by a gauge glass at one corner of it. Along¬ 
side of this gauge glass there is a standpipe from the bottom of 
which the oil flows to the crank chamber. The opening at the 
bottom of the standpipe is controlled by a spring pressed poppet 
valve which is normally held open by the filler cap at the top of 
the standpipe and closes automatically when the cap is removed, 
so no oil will run out. This system is employed on E-M-F motors 
and is illustrated in Fig. 191. 

Automatic Circulation in Crank Chamber—Instead of 
providing a separate feed to each compartment of the crank 
chamber, some manufacturers endeavor to keep the splash level 
in the different compartments substantially constant by means 
of a system of internal circulation. Oil from a reservoir out¬ 
side the engine is pumped through a sight feed on the dash into 
one of the crank case compartments, say the third from the front 
end. It is splashed around by the crank in this compartment, and 
the spray draining back down the walls of the compartment is 
caught in inclined troughs on the walls and conducted into the 
compartment directly in front. There it is splashed over the 
walls again, and again collected in an inclined trough and con¬ 
veyed to the compartment in front. From the foremost com¬ 
partment of the crank case it is conducted into the cam gear 
chamber, and when the oil in the latter reaches a certain height it 
flows through an inclined conduit to the rearmost crank chamber 
•compartment, to be recirculated through the different compart¬ 
ments. The pump simply supplies the loss of oil and does not 
circulate it. 

Manufacturers of two cycle motors drawing their fuel charge 
through the crank chamber sometimes lubricate them by mixing 
oil with the fuel in the proportion of 1 :20. 

Gear Oil Pumps—The most extensively used type of oil 
circulating pump is the gear pump. This consists of a casing in 
which fit snugly two spur gears, one driven by means of its shaft 
and the other by meshing with the former. The oil enters the 
housing on the side on which the meshing teeth separate and fills 
the spaces between adjacent teeth and the wall of the housing. 
It is thus carried around with the teeth to the opposite side of 
the housing and leaves through an opening there. 

A good rule in regard to rate of circulation in circulating lubri¬ 
cation systems is to provide for the delivery per minute of 25 
cubic inches of oil per square inch of projected bearing surface 
supplied directly, at the maximum speed of the motor. If the 
motor is provided with ball bearings and has splash lubrication 


CRANK CASE AND OILING SYSTEM. 


299 


the oil may be circulated at the rate of 15 cubic inches per minute 
per horse power at normal engine speed. Suppose a 4x5 inch 
motor having a i $4 inch crankshaft with three main bearings of 
an aggregate length of inches. The total projected bearing 
surface then is 

9*4 x = 15 square inches, 
and the pump should move 

15 x 25 = 375 cubic inches per minute 
at a speed of, say, 1,200 r. p. m. 

The delivery of oil by a gear pump may be calculated as fol¬ 
lows : Let d be the pitch diameter of one of the gears, p the 

diametral pitch, h the height of the teeth (= - ), f the face width 

P 

of the gears, and n the number of revolutions per minute (which 
is generally equal to one-half the revolutions of the crankshaft). 
It may be assumed that the spaces between the teeth are equal in 
volume to the teeth. The total volume of the annulus comprising 
the teeth and the spaces is 

[V+/A 2 — (d— £) 2 J -j/= irdhf. 

The amount of oil transferred from one side of the housing 
to the other by one gear in one revolution is evidently half this, 
and the two gears together transfer a quantity of oil which is 
represented by the above expression. This is the theoretical de¬ 
livery. The actual delivery is somewhat less, owing to leakage 
over the ends of the gears and between the meshing teeth, but 
in calculating the required size of pump the theoretical delivery 
can be used as a basis. 


Substituting - for h and inserting n the number of revolutions 
P . ’ 

of the pump gears per minute, we have for the quantity of oil 
delivered: 

Q= 2 r d f. n cubic inches -per minute .(93) 

P 

It will be noticed that the delivery is inversely proportional to 
the pitch number, hence a relatively small pitch number should 
be chosen. 

If for the pump for our 4x5 inch motor we choose a pitch of 
8 and a pitch diameter of 1 Y\ inches (10 teeth gears), then at 
600 revolutions per minute the oil delivery will be 

2 v 3.14 x ij x 600 f 

8 

and since 375 cubic inches is required we have 
589 f = 375 





300 


CRANK CASE AND OILING SYSTEM. 


/ = 375 __ 0.635, sa y I inch. 

589 

Feeding a liberal amount of oil to the bearings is desirable, be¬ 
cause the oil, in addition to acting as a lubricant, serves to carry 
off the heat generated by friction in the bearing, and thus tends 
to prevent overheating. 

A design of gear pump for circulating the oil is shown in Fig. 
192. The pump consists of the housing, the cover and the two 
gears. The shaft through which the pump receives its motion 
extends through its bearing, and leakage of oil is prevented by 
means of a stuffing box. The idling gear revolves on a stud se¬ 
cured into a blind hub formed on the housing. The inlet and 
outlet openings are on opposite sides. Gear pumps have compara¬ 
tively little suction power, and for this reason are generally 



placed at the side of the oil well so that the oil flows into them 
by gravity. 

German designers often provide double gear pumps, there being 
one set of gears on each side of the bearings. One set of gears 
is of about half the width of face as the other, the smaller pump 
delivering oil to the circulation indicator and the larger one to 
the distributor. The arrangement has the advantage that the 
bearings are nearly symmetrically loaded, but its complication is 
against it, and the indications of the circulation indicator are 
hardly as reliable as when the oil delivered by the pump which 
supplies the bearings flows through the indicator. 

Sliding Vane Pump—This type of pump (Fig. 193), which 
is used to a small extent for circulating the lubricating oil, con- 






























































301 


CRANK CASE AND OILING SYSTEM. 



sfsts of a cylindrical chamber in which is located a disc ot a 
thicimess equal to the internal height of the chamber but of 
smaller diameter. The disc is located eccentrically with respect 
to the chamber, and is cut with a diametral slot dividing it into 
halves, in which slot are located two sliding vanes which are 
pressed apart by a flat spring between them. It is apparent that 
the volume of liquid moved by such a pump per revolution is 

substantially equal - 

to the difference be- - 

tween the volume 
of the pump chamber and that 
of the disc within it. The hous¬ 
ings of gear and sliding vane oil 
pumps are made of brass, and the 
rotating parts of steel, but the 
gears may be of cast iron. 

Plunger Pump—Of late 
plunger pumps have come into ex¬ 
tensive use for circulating the oil, 
being generally placed inside the 
crank chamber and operated 
directly from the camshaft. They 
may be (and occasionally are) 
actuated by one of the valve 
cams, but it is better to employ 
an eccentric on the camshaft for 
the purpose, which gives a 
smoother action. Referring to 
Fig. 194, A is the 
pump barrel, made 
of brass tubing, 
which is inserted 
into the crank cham¬ 
ber from below 
and is held by means 
of rings cast on the 
crank chamber wall. Fig. 194.— Plunger Oil Pump in 
Within the barrel A Crank Chamber. 

is the plunger B, 

which is drilled out at its lower end to receive the coiled spring 
C. At the bottom of the pump barrel is located a ball valve D. 
The spring rests upon the valve housing and forces the plunger 
against the eccentric E on the camshaft. As the plunger moves 
upward the space in which the spring is located increases in 

















































302 


CRANK CASE AND OILING SYSTEM. 


volume and oil enters it through the ball valve. When the 
plunger is next moved down by the eccentric the ball valve 
closes and oil is forced through the drill hole F in the plunger, 
the space formed by the flat on the plunger, and out through the 
check valve G at the delivery opening. The delivery per stroke 
is, of course, equal to the change in volume of the spring 
chamber. 

Oil Strainers—The oil drawn in by the pump must be 
strained, so that any metal dust worn off the bearings may be 
separated from it, because if this metal dust should get into the 
bearings it would start cutting. The strainer should be made of 
brass wire gauze of about 16 meshes to the inch, of such form 
that it can be easily removed for cleaning. It is also well to 
remember that as the oil becomes dirty it will clog up some of 
the meshes, and in order that the flow of oil through the strainer 



Overflow Pipe and Drain. Oil Strainer. 

may not be appreciably impeded it should be made of large area. 
One design of combined drain, strainer and overflow pipe is 
shown in Fig. 195. This is in the form of a hollow cylindrical 
casing screwing into the bottom wall of the crank chamber, 
which is covered with a layer of wire gauze. The joint in the 
bottom of the oil well is rendered oiltight by means of a gasket 
under the flange of the strainer. 

Frequently the strainer is made funnel shaped, with the edge 
of the big end flared outwardly, and is inserted into the oil 
well through the opening over which the pump is bolted. Still 
another design of strainer, as used on the Speedwell motor, is 
shown in Fig. 196. In this case a gauze drum extends horizon¬ 
tally through the oil well, one end fitting against a tapered seat 
at the outlet to the pump and the other end being pressed 












































CRANK CASE AND OILING SYSTEM. 


303 


against by a large plug closing the opening through which the 
strainer is inserted. 

Oil Level Gauge—When the supply of oil is carried in an 
oil well at the bottom of the crank case some means must be 
provided to enable the operator to determine quickly how much 
oil remains in the well. Some manufacturers merely provide 
two try cocks in the side of the oil reservoir, one of which is- 
placed at the lowest level 
which the oil should be allowed 
to reach. Others place a “bull's- 
eye” or glass disc in the side of 
the well. This is not very satis¬ 
factory, since the oil quickly as- 
cumulates dirt, which settles on 
the glass and makes it practically 
impossible to distinguish the oil 
level. 

The most satisfactory form of 
level gauge is a float in a float 
chamber communicating with the 
oil well, as shown in Fig. 197. 

The float is provided with a 
stem, which extends up into a 
glass tube at the side of the 
upper half of the crank cham¬ 
ber, with a black ball on top, 
which, in connection with scale 
marks on the glass, gives an ac¬ 
curate idea of the height of the 
oil in the well. Instead of bolt¬ 
ing the float chamber to the oil 
well it is sometimes cast integral 
with the lower half of the crank 
case, while again the float 
may be placed in the oil well 
itself. 

Circulation Indicator A Fig. 197.— Oil Level Indicator. 
continuous supply of lubricant to 

its bearing surfaces is almost as essential to the proper opera¬ 
tion of a gasoline motor as breathing is to the sustenance of 
human life. If the oil feed stops it is usually only a short 
time until destructive heating and cutting of the bearings begins,, 
and there should, therefore, be some device constantly indicat¬ 
ing to the driver that the oil is being fed properly. If oil is- 



__ lji lL Fjppl 









































304 


CRANK CASE AND OILING SYSTEM. 


supplied to the bearings under pressure a pressure gauge or in¬ 
dicator on the dash will serve the purpose, since the oil feed 
could only stop if the pump should cease to work or an oil 
lead should break, in either of which cases the pressure in the 
system would become nil. 

In the case of the ordinary circulating systems the general 
practice is to carry either all the oil or that fed to one bearing 
or one crank case compartment, to the dashboard, where it 

passes through a circulation indicator. 
The most familiar form of circula¬ 
tion indicator is of the fountain 
type, as illustrated in Fig. 198. It 
consists of a brass base bolted to 
the dash, in which there are two 
oil channels with threaded connec¬ 
tions in front of the dash. The oil 
channels open on the top side of the 
base, and one, the inlet channel, is 
provided with a short standpipe, the 
top end of which is curved in a half 
circle, so that it discharges down¬ 
wardly. On the base rests a glass 
cylinder with a brass cap which is 
screwed to a threaded projection of 
the standpipe at the centre of the 
glass cylinder. When the oil is cir¬ 
culating it gushes from the stand¬ 
pipe. 

Oil Distributor —The oil dis¬ 
tributor used in some motor lubri¬ 
cating systems, and which is some¬ 
times combined with the pump, con¬ 
sists of a housing in which rotates 
a disc with an oil channel in one 
face which places a central opening 
in the cover through which the oil 
arrives, in communication successively 
with delivery openings arranged concentrically. Fig. 199 shows 
the construction. The oil ports should be so arranged that the 
passage through the distributor is never closed. If some of the 
bearings supplied by the distributor require more oil than others 
this can easily be arranged for by making the outlet ports so as 
to extend over different angles. The oil distributor must be 
driven from the engine, and would generally be connected to 
the rear end of one of the camshafts. 




Fig. 198.— Circulation 
Indicator. 
































CRANK CASE AND OILING SYSTEM. 


305 



Fig. 199.—Oil Distributor. 


Filling Holes and Drain Plugs—Oil is generally poured 
into the crank chamber through the breather tubes. The or¬ 
dinary design of breather tube, however, does not admit of 
pouring the oil in rapidly, and for this reason breather tubes 
are now sometimes made funnel shaped with a hinged cover, as 
shown in Fig. 200. The cover is provided with a wire screen 
through which the motor “breathes.” 

Drain plugs or cocks should be placed in the bottom of the 
oil well so as to permit of draining off all oil after it has be¬ 
come contaminated. Provisions should also be made for drain¬ 
ing the splash chambers, and it is a good plan to have the bot¬ 
tom of the oil well and splash chambers slope slightly toward 
the drain plug. All cocks fitted to the bottom of automobile 
motors should be so arranged that they are closed when the 
handle points downward, so they 
cannot possibly jar open. 

Size of Oil Conduits—The oil 
pipes and conduits in circulating 
systems are frequently made of 
inadequate size, which is probably 
due to the fact that in systems 
where oil is fed as it is consumed 
the rate of feed is very slow, and L 
small tubes are therefore 


very small tubes are 
used; so engineers have come to 
consider very small tubes adequate 
for conveying oil. In circulating 
systems the rate of flow through 
the tubes, based on the theoret¬ 
ical delivery of the pump, should 
not exceed 200 feet per minute, 



Fig. 200.—Oil Filler 
Hole and Cap. 





















































306 


CRANK CASE AND OILING SYSTEM. 



Milling Machine. 




















CRANK CASE AND OILING SYSTEM. 


307 


and conduits formed in the crank case should preferably be 
larger. All oil holes from pockets to bearings should also be 
of large size, because there is no pressure on the oil caught in 
the pockets, and when it is cold it does not flow easily through 
small holes. 

Manufacture of Crank Cases—One of the first operations 
on the castings of a divided crank case is to face off the surfaces 
for the joints between the crank case and the cylinders and be¬ 
tween the several parts of the crank case. Owing to the way 
these surfaces are spread out, particularly the surfaces at the joint 
between the halves, and to the fact that the castings are of alu¬ 
minum, an easy cutting metal, these operations are preferably per¬ 
formed m a milling machine. Fig. 201 shows a crank case being 
machined in a Cincinnati vertical milling machine. The cutter is 
15 inches in diameter and runs at 224 r. p. m., using a feed of 
20 inches per minute. From inch to inch of metal is re¬ 
moved in the roughing cut, and the finishing cut is about 0.010 
inch deep. The work is done by first feeding across the piece to 
rough it off, then reversing the feed and coming back at the 
same rate for the finishing cut. The table is, of course, raised the 
0.010 inch after the roughing cut has been taken and before the 
machine is reversed for the finishing cut. The surfaces produced 
are flat under a straight edge within 0.001 inch and parallel within 
the same limit. 

In crank cases with tunnels for the camshafts there are three 
boring operations to be performed; namely, the boring for the 
crankshaft bearings or bearing plates and the borings of the two 
camshaft tunnels. These three operations are preferably per¬ 
formed at the same time, which can be done in a horizontal bor¬ 
ing mill with special fixtures, as shown in Fig. 202. 

A crank case requires numerous holes to be drilled, and this is 
one of the points where labor-saving machine tools can be applied 
to advantage. The top surface of the top half requires holes to be 
drilled for the cylinder bolts, the push rod guide bolts, and possi¬ 
bly also oil holes to the crankshaft bearings. The bottom surface of 
the top half requires holes to be drilled for the bolts holding the 
upper and lower halves together, for the bearing cap bolts, the 
bearing bushing dowels, and for oil passages from oil pockets. 
The lower half also requires numerous holes to be drilled. For 
all these operations a multiple drill press is a very serviceable tool. 
Fig. 203 shows a Baush multiple drill installed in the Cadillac 
plant in Detroit, set for drilling all of the drill holes in the top 
of a barrel type crank case. As many as twenty-two holes are 
thus drilled at the same time, at two or three different levels. 


308 


CRANK CASE AND OILING SYSTEM 



ig. 202.—Boring Crank Case by Means of Three-Spindle Jig in Bement 46-lNCH Horizontal Boring Machine. 













11 ?1 


CRANK CASE AND OILING SYSTEM 


309 



Fig. 203.—Drilling Crank Case in Baush Multiple Spindle. 

Drill. 
















310 CRANK CASE AND OILING SYSTEM. 

A drilling jig is used, and the holes therefore must come ac¬ 
curate. 

Bearing Bushings—There are two types of bearing bushings 
used in crank cases, namely, integral bushings and split bushings. 
The former are produced by standard shop processes and require 
no special equipment. Manufacturers of bearing metals now 
furnish these metals in cylinders of many different sizes, from 
which the bushings may be turned up directly. The manufac¬ 
ture of split bushings, on the other hand, involves some difficul¬ 
ties, and special jigs can be used to advantage. There are really 
three types of these bushings, viz., bronze bushings, babbitted 
bronze bushings and babbitt bushings. These will be consid¬ 
ered in rotation. 

Manufacture of Split Bushings—The old method of making 
split bronze bushings consists in casting the bearings solid in 
halves, planing the faces which are to go together and then 
soldering the halves together, after which they can be ma¬ 
chined like single piece bushings. After the machining opera¬ 
tions are completed the bushings are heated to such a degree 
that the solder melts, and the halves are then separated and the 
solder is wiped off. This method works very well on a small 
scale, but, of course, some of the pieces are almost certain to 
come apart in the turning, and, besides, it is difficult to get the 
two halves absolutely identical by the ordinary methods of 
machining. 

The jigs shown in Figs. 204 and 205 were designed by Herbert 
E. Barnes for machining split camshaft bushings that must be 
parted accurately in the centre, and were described by him in 
The Horseless Age of January 29, 1908. The bushings are cast 
in halves, with only little stock to remove on the inside. First 
they are milled off on the faces which come together, and then 
they are clamped singly in the jig, Fig. 204, to have the ends 
faced off and chamfered on the outside. The object of this 
chamfering is to permit of accurately centreing the bushing in 
the drilling jig, Fig. 205, and holding it on the cup mandrel, 
Fig. 206, for turning the outside. The facing and chamfering 
jig, Fig. 204, consists of a cast iron frame A, a table B held to 
the frame by screws and dowels, machine steel clamps C C, be¬ 
tween which and the table B the bushing is clamped, and shoulder 
bushings D D to guide the tool L. The latter is made from tool 
steel, with milling teeth on the end extending over a little more 
than half the circumference. These teeth serve for facing the ends 
of the bushings, and the depth of cut is governed by the adjust- 


CRANK CASE AND OILING SYSTEM. 


311 


able locknuts M M, which are brought up against the ends of 
bushings D D. In the side of the tool is a slot into which the 
chamfering, blade N is inserted. 

The jig, Fig. 205, for drilling and reaming the hole in the 
bushing comprises a suitable cast iron frame for holding two 
cup bushings, one, A, being fixed for the reamer, and the other, 
B, being threaded to adjust up and down in a tapped hole, for 



Fig. 204. Jig for Facing and Chamfering Ends of Half 

Bushing. 


the drill. The hole in the drill bushing is counterbored at the 
inner end, so that the reamer will not be injured when it goes 
through the work. 

When the two mates are put in this jig and the screw bushing 
B is adjusted they are held by the floating clamp C and the strap 
on the swinging screw F. The clamp C rests on the two bosses 
















































































































































































312 


CRANK CASE AND OILING SYSTEM. 


G, ,G, and is controlled by two studs H, H, which fit in the clamp 
and also extend through the bosser G, G. The holes in these 





Fig. 205.—Jig for Drilling and Reaming Bushings and Reamer. 

bosses are one-eighth of an inch larger than the studs, which in 
turn are at least one-sixteenth of an inch larger than will permit 
them to bind the clamp. 










































































































































CRANK CASE AND OILING SYSTEM. 


313 


The clamp Y has teeth milled in its recess to grip the casting 
and prevent its turning with the tool, without undue pressure 
being used to hold it. This freedom of the clamp is provided so 
as to avoid any tendency to cramp the work or displace it in the 
cup bushings. 

A four lip, straight flute drill, ground nearly flat, will give the 
best results because of the cored hole in the bushing which is 
t>eing machined. The reamer, shown in Fig. 205, should have its 
flutes cut in left hand spirals, and should be so made that the 
body is long enough to guide it all the way through the work 
that it does not leave the bushing during the entire cut. The 
cutting blades need not be more than 1^4 inches long. 

Fig. 206 shows a mandrel for turning the outside of the bush¬ 
ings. It consists of a piece of steel turned to the proper size 
with a solid cup and a movable cupped collar, which is tightened 
by the nut and holds the work during the turning. 



Fig. 206.—Mandrel for Turning Split Bushings. 


Babbitt Lined Bronze Bushing—A somewhat similar method 
for producing split crankshaft bushings was described by O. 
Linley in the Automobile Engineer (London) for February, 1911. 
It is claimed to make the half bushings interchangeable, so it is- 
not necessary to keep them together in pairs. 

There is practically no machining allowance on the joint, the 
first operation being to face the halves up on a disc grinder. 
When this is done each half is taken singly and attached to the 
jig A, Fig. 207. This jig consists of a piece of mild steel turned 
with a wide collar, which is subsequently half removed by mill¬ 
ing, and the half bushing it attached by means of the clip shown 
at B. Eye and hand alone are relied on for centreing the bush¬ 
ing longitudinally, so the attachment is a very quick and easy 
job, and the flanges can be turned to gauge size with almost 
equal thickness. The next operation is to place a pair of bush¬ 
ings together in a split chuck for boring, turning the radius on 
the outer end, and recessing by the tool shown at C, the inside 
being left rough so as to provide a good surface to take the 
babbitt metal. The bushings are then babbitted and are bored, 

































314 


CRANK CASE AND OILING SYSTEM. 


out to size, the radius being put on at the end at the same time, 
and the outside is turned on the cup arbor shown at D. 

Manufacture of Split Babbitt Bushings—Babbitt bushings 
are cast integral in the first place, generally & inch thick, with 
an inside diameter 0.03 inch small. It is advisable to consoli¬ 
date or “compact” the babbitt so as to render it capable of 
withstanding greater bearing pressure, and to this end the cast 
bushing is placed in a die which is provided with several keyt 
slots, and a straight and a tapered mandrel are then successively 
driven through the bushings, which “compacts” the metal and 



Fig. 207.— Tools for Making Interchangeable Half Bushings. 

forces some of it into the keyways. These mandrils leave the 
bushing very slightly under size (from 0.001 to 0.003 inch). The 
remainder of the stock on the inside diameter is preferably re¬ 
moved by a reamer when the bushing is in place. The bushing 
is next split diametrically by means of a one-thirty-second inch 
saw, which necessitates the use of shims between the halves to the 
thickness of 0.033 inch. The bushings for the crankshaft main 




















































































CRANK CASE AND OILING SYSTEM. 


315 


bearings are reamed in place by means of a long reamer in a 
horizontal or vertical boring machine, the bearing being left, say, 
o.ooi inch large. The keys on the upper halves of the crank¬ 
shaft bushings for an engine in which all bearings are supported 
by the upper halves of the crank case should preferably be 
smoothened off some, so that these halves may be removed from 
the crank case without entirely removing the crankshaft, by 
“rolling” them around the shaft. 

It has been found that in order to reduce the friction of a 
bearing to a minimum the bearing bushing must be relieved at 
the sides so that the journal is in contact with the bushing only 
on top and below. Rules for calculating the reduction of the 
friction and the increase in the unit pressure for different reliefs 
are given by Goodman in his “Mechanics for Engineers.” In 
order to relieve the bushings the bearing caps are loosened and 
shims to the proper thickness placed between the upper and lower 
halves, whereupon reamers of gradually increasing size are run 
through the bearings until the full relief is attained. 


CHAPTER XIV. 


STARTING CRANK, MANIFOLDS AND MUFFLER. 

Starting Cranks—The ratchets of starting cranks are made 
in many different forms, several of which are illustrated in Fig. 
207. That shown at B is formed on the shaft of the starting 
crank itself, which is drilled out for the purpose, and has five 
end ratchet teeth milled on it. Frequently the engaging faces 
of these teeth are in a plane passing through the axis of the 
shaft. It is a much better plan, however, to make them at an 
angle of about 10 degrees with that axis, as in designs A, B 
and D, so that the teeth of the two ratchets will interlock, for 



Fig. 207.—Forms of Starting Crank Ratchets. 


even though these ratchets are hardened, their teeth will wear 
in time, and when the points of the teeth have worn back the 
ratchet is liable to disengage while the operator is pulling hard 
on the starting crank, which is extremely annoying if not dan¬ 
gerous. It may here be pointed out that practically all motors 
turn right-handedly when looked at from in front of the car 
or the starting end. 




316 















































































STARTING CRANK, MANIFOLDS AND MUFFLER. 317 

Starting crank ratchets for 
four cylinder motors should 
preferably have two teeth, 
spaced 180 degrees, so that the 
starting crank handle will al¬ 
ways be in the same position 
when the crank shaft passes 
the dead centre. For the same 
reason the ratchet for six 
cylinder starting cranks should 
have three ratchet teeth, spaced 
120 degrees, as shown at D. 

In some cases the ratchet on 
the end of the crankshaft is 
replaced by a pin passing 
through same, with projecting 
ends with which the teeth on 
the starting crank ratchet en¬ 
gage, the teeth then being made 
of special form, as shown at 
C, Fig. 207. Generally the 
ratchet on the forward end of 
the crankshaft is keyed or 
pinned to the shaft, but some¬ 
times, in order to save space 
in the longitudinal direction, 
it is formed integral with the 
crankshaft pinion (see A, Fig. 207). 

Angular Relation of Starting Crank to Crankshaft—In 

order to render the starting of the motor as easy and safe as pos¬ 
sible the ratchets on the starting crank and the crankshaft must 
occupy certain definite angular positions. The resistance which 
the operator must overcome in turning the motor over by hand is 
due partly to the friction of the pistons in the cylinders and partly 
to the compression in one cylinder, the latter being generally the 
most important factor. The compression reaches its maximum 
value when the piston is at the top end of the stroke, but the 
crank effort necessary at that moment is zero, owing to the tog¬ 
gle effect of the crank and connecting rod. The crank effort re¬ 
quired at various angles of the crankshaft during the compres¬ 
sion stroke is shown by the diagram, Fig. 208, from which it 
will be seen that this effort reaches its maximum when the crank 
is about 25 degrees ahead of dead centre. 

The usual method of starting the motor consists in pulling up 



Fig. 208.—Effort Required on 
Starting Crank at Dif¬ 
ferent Crankshaft 
Angles. 








318 STARTING CRANK, MANIFOLDS AND MUFFLER. 


on the starting crank, and the operator undoubtedly applies his 
strength to the best advantage when the starting crank is in the 
horizontal position or half way up. Hence the ratchets on the 
starting crank and the crankshaft should be so arranged that the 
starting crank is half way up, or 90 degrees from the straight-up 
position, when one set of cranks has to travel 25 degrees more 
till it reaches the top dead centre position. A four cylinder crank¬ 
shaft usually stops in such a position that all of its pistons are 
about mid-stroke, hence the crankshaft must be turned through 



Position of Crankshaft 
and Starting Crank. 


a quarter circle to bring it to the 
dead centre or firing position, 
and since when the crankshaft 
is in the dead centre position 
the starting crank will be 90 — 
25 = 65 degrees from the straight- 
up position, it will generally 
be possible to pick up the crank¬ 
shaft at 65 + 90=155 degrees 
from the straight-up or 25 de¬ 
grees from the straight-down 
position. 

The turning moment which it 
is necessary to impress upon the 
starting crank is proportional to 
the square of the bore and to the 
length of the stroke. It is, of 
course, impossible to make the 
length of the starting crank pro¬ 
portional to the turning moment 
required, as that would give ex¬ 
ceedingly ’small cranks for small 
motors and entirely too large 
cranks for large motors. The fol¬ 
lowing formula gives suitable val¬ 
ues for the length of the crank: 


Lc — d V / .(74) 

The necessary cross-section of the crank arm may be calculated 
on the basis of 12,000 pounds per square inch stress in the metal 
when the operator applies an effort of 100 pounds to the crank 
handle. Supposing the starting crank to be 10 inches long, the 
bending moment is 

10 x 100 = pounds-inches. 

The section of the crank must, therefore, have a resisting moment 



















STARTING CRANK, MANIFOLDS AND MUFFLER. 319 

of 1,000 pounds-inches when subjected to a fibre stress of 12,000 
pounds per square inch. In the case of the 4X5 inch motor used 
in our calculation, making the thickness or smallest dimension 
of the arm one-half inch, the width or largest dimension near the 
hub figures out to about one inch, if the arm is made of solid 
rectangular cross-section. The crank arm is frequently drop 
forged with a channel section, the open side of which is placed 
toward the rear. The starting handle should be about one inch 
in diameter and not less than 5X inches long. It should be pro¬ 
vided with a brass tubular or hardwood grip which turns freely 
on the pin or stud, so as not to injure the operator’s hands. The 
handle is sometimes drilled out for lightness and is generally ex¬ 
panded into the crank arm. 



Fig. 210.—Starting Crank Supporting Clip and Swinging 


Bracket. 

Locking Starting Crank —The starting crank must not be 
allowed to hang down when the car is in motion, as its grip is 
generally considerably lower than the front axle, and is therefore 
likely to be struck by obstructions on the road. A leather socket 
can be slipped over the grip, suspended by means of a strap from 
the chassis frame in such a way that the crank rests in a nearly 
horizontal position. 

On motor trucks it is not desirable that the handle of the 
steering crank extend forward when not in use. For convenience 
in starting the handle must project ahead of the bumper with 
which most trucks are fitted, and if it does this while not in use 
it is apt to be damaged by collision with some object. To pre¬ 
vent such damage the shaft of the crank on motor trucks is gen¬ 
erally provided with a hinge joint so that after the motor has 
been started the crank can be swung back and held in a clip- 
secured to the front cross member of the frame. Fig. 210 shows 
one construction serving this purpose. Instead of the shaft of 










































•320 STARTING CRANK, MANIFOLDS AND MUFFLER. 


the crank being hinged it is supported in a rotatable bearing. 
The sub-figure shows a simple form of clip for this purpose. 

Fig. 211 shows the variation of the starting and low speed run¬ 
ning resistance of automobile motors as a function of the total 
cylinder displacement in cubic inches. This diagram, which is 
taken from a paper read by C. E. Wilson before the Society of 
Automobile Engineers, is of value in the design of self-starters. 

Inlet Manifolds. —The inlet manifolds present a special prob¬ 
lem for the designer, since they must be so designed that each 
cylinder of a multi-cylinder motor receives substantially the same 



Fig. 21 i.—Starting and Running Torques of Four and Six 

Cylinder Motors. 


charge, both quantitatively and qualitatively. That is, each cylinder 
must receive the same amount of combustible mixture as every 
other cylinder, and the mixture received by any cylinder should 
contain the same proportion of gasoline as that received by any 
other cylinder, thoroughly gasified and mixed with the air. 

Where the cylinders are cast in pairs the inlet ports are now 
almost universally siamesed, so that in the case of a four cylin¬ 
der motor the “manifold” has only two branches and is really a 
\ or a T. Inlet manifolds are cast of aluminum, brass or malle¬ 
able iron, and are sometimes made of drawn brass tubing and 
•cast fittings or solely of brass tubing bent and brazed together. 







































STARTING CRANK, MANIFOLDS AND MUFFLER. 321 

1 he inlet pipe is made with an internal diameter equal to 36 to 40 
per cent, of the cylinder bore. Its length varies according to the 
location of the carburetor. When the fuel is fed by gravity from 
a tank located inside the front seat the carburetor must be located 
low, so that the fuel may always be under a head, even on 
the steepest upgrades. On the other hand, when the fuel is fed 
by gaseous pressure the carburetor is placed as high as possible, 
in order that its adjusting means may be as accessible as possible. 
Sometimes the carburetor is located on the opposite side of the 
motor from the valves, and the inlet pipe must then extend across 
the top of the engine or a cored inlet passage must be provided 
in the cylinder casting from the valve side to the opposite side. 
On the whole a short length of exposed inlet pipe is preferable, 
because in cold weather the gasoline, which has been vaporized 
in the carburetor by means of the heat supplied by the water 



circulating through the jacket, will tend to condense again in the 
inlet pipe. 

Three different types of inlet manifolds for four cylinder motors 
with siamesed inlet ports are shown in Fig. 212. They are of T, 
Y and V form respectively. Any intermediate form may be used, 
and the manifold may be curved in various ways to pass around 
obstacles; but it is advisable to reduce the bends to the smallest 
possible number and to make those that cannot be eliminated of 
as large a radius of curvature as possible, since the curves greatly 
increase the resistance to gaseous flow. Some designers con¬ 
siderably enlarge the cross-section where the branches join, so 
as*to avoid the formation of eddies at these points. Manifolds 
with curvatures in more than one plane are sometimes made in 
two parts bolted together, so as to avoid difficulties in casting. 














322 STARTING CRANK, MANIFOLDS AND MUFFLER. 



Fig. 213.—Single Bolt Manifold Joint. 


Manifold Joints—The general practice is to secure the inlet 
manifold to the cylinder casting and to the carburetor by means 
ot flanges and cap screws, 1 % inch screws being used. If the inlet 
and exhaust manifolds are both located on the same side of the 
motor it is somewhat difficult to get at screws through lugs, and 
the two manifolds are then generally held on by means of yokes 
and studs. Still another method of fastening the inlet manifold 
to the cylinder casting consists in screwing a stud into the cylin¬ 
der wall at the centre of the inlet port, this stud extending 
through the wall of the manifold, which is clamped in place by 
means of a nut on the stud, as shown in Fig. 213. 

The lower end of the inlet manifold connects to the carburetor 
by means of a flanged joint. These joints have been standardized 
by the Society of Automobile Engineers, the dimensions of the 
standard flanges being given in Table VI. 

TABLE VI—CARBURETOR FLANGES. 



Carburetor, 


Size. 

A. 

B. 

C. 

D. 

E. 



P\ 

H 

II 

11 1 ? 

1 H 

11 

i 9 e xi8 U. 

S. 

s. 

H 

Vs 

1 -fa 

iVs 

1 y* 

11 

&xi8 U. 

S. 

s. 

H 

1 

1 1 3 ? 

1 i *8 

iVs 

11 

i 5 b xi8 U. 

s. 

s. 

11 


x tV 

ill 

2 fs 

11 

Hxi6 U. 

s. 

s. 

•11 


ill 

ill 

2 'A 

11 

Hxi 6 U. 

s. 

s. 

11 


ill 

ill 

211 

11 

IB x 14 U. 

s. 

s. 


2 

2 A 

ill 

3 Vs 

11 

iVxi 4 U. 

s. 

s. 

■ft? 














































STARTING CRANK, MANIFOLDS AND MUFFLER. 323 

Inlet manifolds for six cylinder motors with siamesed inlet 
ports may be made in the form of a T, but are more frequently 
made in the form shown in Fig. 214, in which design the distance 
from the carburetor to each of the inlets ports is substantially the 
same, which tends toward more uniform distribution. It will be 
noticed that the intermediate distributing branches join the final 
distributing pipe somewhat closer to the central inlet port than to 
the outer ones. 

Exhaust Manifolds—In T-head motors the exhaust ports are 
also generally siamesed, though, as already explained, if it is de¬ 
sired to get the most possible power out of the motor this should 
not be done. On L-type motors and motors with valves in the 
head the exhaust ports are separate, and one connection to the 
manifold is then required for each cylinder. The exhaust mani¬ 
fold is generally made of an internal diameter equal to one-half 
the cylinder bore. It is made of malleable iron or cast steel anc 
occasionally of steel tubing. Frequently the exhaust manifold i« 



cast with four longitudinal ribs to help cool it. The walls are 
generally cast one-eighth of an inch thick. The exhaust manifold 
is fastened to the cylinder casting in the same manner as the 
inlet manifold, and at its rear end it turns down through an 
angle of 45 degrees and is flanged for connection to the exhaust 
pipe leading to the muffler. Some designers increase the diam¬ 
eter of the manifold from the front to the rear end, but usually 
it is made of uniform size throughout its length. In six cylinder 
motors the exhaust manifold is sometimes in telescoping sections, 
permitting of its free longitudinal expansion and contraction. The 
manifold may reach almost red heat when the motor operates 
continuously at full load, and will naturally expand considerably. 
The crank chamber, on the other hand, which determines the 
distance apart of the cylinders, does not appreciably heat up; 
hence if no provision was made for the expansion of the mani¬ 
fold it and the cylinders would be severely strained. There is, of 











324 STARTING CRANK, MANIFOLDS AND MUFFLER. 

course, the same need for expansion joints in a four cylinder ex¬ 
haust manifold as in a six cylinder. 

Fig. 215 shows an arrangement of inlet and exhaust manifolds 
on the same side of the motor. The exhaust manifold is then 
generally swung upward and the inlet manifold downward so as 
to avoid interference, and the two are mostly secured in place 
by means of yokes. 

Effect of Timing Order—The explosion in a four cylinder 
motor may be timed in either of two orders, viz.: 

1—2—4—3, 

1—3—4—2. 

With siamesed inlet ports and a symmetrically arranged inlet 
manifold neither of these firing orders has any appreciable advan¬ 
tage over the other. But if there are four individual cylinders and 
a manifold of T form is used, then, it is claimed, the last men- 



Fig. 215. —Inlet and Exhaust Manifolds on Same Side of 
Motor and Secured by Yokes. 

tioned firing order is the preferable one, for the following reason; 
it will be noticed that with the first mentioned firing order the 
inner cylinder of a pair draws in its charge immediately after its 
mate. It was pointed out in the chapter on “Valves and Valve 
Gear” that the inlet valve is generally given a lag of 30 degrees, 
so as to enable it to fill more nearly completely. The inlet valve 
opens with a lag of only 10 to 15 degrees. Now, suppose cylin¬ 
der No. 1 to have completed its inlet stroke. About 10 degrees 
later, while its inlet valve is still being held open to enable it to 
draw additional charge, the inlet valve of cylinder No. 2 opens, 
and since there is considerable suction in this cylinder and it is 
nearest to the carburetor, the flow of gas to cylinder No. 1 is 
thereby prematurely stopped. On the other hand, with the tim¬ 
ing order last mentioned the outer one of a pair of cylinders 
always follows the inner one, and since the distance from the 
carburetor to the outer cylinder is greater than to the inner cyl- 

























Fig. 217.—Manograph Diagram Showing Effect of Exhaust 
from one Cylinder Entering Another. 


STARTING CRANK, MANIFOLDS AND MUFFLER 325 


inder, the inner cylinder is 
not “starved” by the open¬ 
ing of the inlet valve to the 
outer cylinder—at least not 
to the same extent. 

The firing order is of 
more importance in a six 
cylinder motor, however. 

We found that the exhaust 
valves remain open for about 
230 degrees of crank motion. 

Now, in a six cylinder motor 
the exhaust valves begin to 
open at intervals of 120 de¬ 
grees, so two exhaust valves 
are open simultaneously prac- 
ticaly all the time. This 
overlapping of the exhaust periods is shown graphically in 
Fig. 216. Suppose that the exhaust valve opens with 45 
degrees lead and closes with 5 degrees lag; then, when the 
exhaust valve of one cylinder begins to open the exhaust valve of 
the preceding cylinder is still fully open, and remains open dur¬ 
ing another no degrees of crank motion, so that if the two cyl¬ 
inders are adjacent to each other and connected to the same ex¬ 
haust manifold, and especially if the exhaust ports of the two 
cylinders are siamesed, exhaust gases must “blow over” from the 


Ftg. 216.—Overlapping of Ex¬ 
haust Periods in Six Cyl¬ 
inder Motors. 










326 STARTING CRANK, MANIFOLDS AND MUFFLER. 

cylinder just beginning to exhaust into the one completing its ex¬ 
haust stroke, thus interfering with the evacuation of the latter. 
This effect is brought out very clearly in the manograph diagram 
shown in Fig. 217, for which the writer is indebted to Hugo Gib¬ 
son, of New York. This diagram was taken from a six cylinder 
motor in which this “blowing over’' effect had not been pro¬ 
vided for. 

This difficulty can be partly overcome by so arranging the fir¬ 
ing order that the two cylinders in one pair never fire in succes¬ 
sion. For instance, the following three firing orders may be used: 

i - *—4 2 6 3 5 > 

1—5—3—6—2—4, 

1 —3—2—6—4—5 • 

The surest way to prevent interference of the exhausts from 
succeeding cylinders is to provide two exhaust manifolds, one for 
the three forward and the other for the three rear cylinders, con¬ 
necting to a double exhaust pipe leading to the muffler. 

Muffler Design—Automobiles must operate substantially noise¬ 
lessly, and the gases from the exhaust must be passed through a 
muffler before they are discharged to the atmosphere. In the 
muffler the gases must be allowed to expand gradually and to 
cool, thereby reducing the pressure, which is the cause of the 
noise when they are discharged into the atmosphere directly. 
While it is an easy matter to so obstruct the passage of the gas 
from the cylinder to the atmosphere that the gases will be dis¬ 
charged without disagreeable noise, they must escape w r ith com¬ 
parative freedom, so that there will be no back pressure from 
them on the piston during the exhaust stroke, and this compli¬ 
cates the problem of muffler design. 

The muffler is generally carried under the rear part of the 
frame of the car, as far away from the motor as possible, so 
the motor will exhaust into an exhaust pipe whose capacity is two 
to' four times that of a single cylinder. This gives the hot gases 
a chance to expand and cool. It is, however, advisable to pro¬ 
vide additional expansion volume in the muffler. The combined 
volume of the exhaust pipe and that compartment of the muf¬ 
fler which the gases enter first should not be less than six times 
the piston displacement of a single cylinder. This expansion 
chamber should have as large a surface as possible in contact 
with the atmosphere, so that heat may be rapidly abstracted from 
the gases. From the expansion chamber the gases are led through 
another chamber or series of chambers communicating by means 
of fine and sometimes tortuous passages, before they are finally 
discharged to the atmosphere. Sometimes these muffling cham- 


STARTING CRANK, MANIFOLDS AND MUFFLER. 327 
• 

bers are placed concentric with the expansion chamber and some¬ 
times they are placed adjacent to it. 

The muffler should preferably be so arranged that it can be 
taken apart for cleaning, since lubricating oil and solid carbon 
particles held in suspension by the exhaust gases will be depos¬ 
ited on the muffler walls, clogging the fine passages and thus in¬ 
creasing the back pressure. The muffler must, however, be of 
sufficiently strong construction to easily withstand the pressure 
of an explosion of a gasoline-air mixture at atmospheric pressure, 
since such explosions in the muffler are of common occurrence. 
The walls of the muffler must be so designed that they will not 
be set into harmonious vibration by the succeeding exhausts, and 
especially should they not be of any material with sonorous 
properties. A covering of sheet asbestos over the muffler will pre¬ 
vent ringing, but is undesirable on account of its heat insulating 
properties. The outer cylindrical shell should preferably be made 



Fig. 218.—Concentric Cylinder Muffler. 


of very thin and soft sheet iron formed into a roll of several 
layers, held together by rivets. F. E. Watts found (The Horse¬ 
less Age of October 2, 1907) that the entire capacity of the muf¬ 
fler is generally about thirteen times the piston displacement in 
one cylinder. 

A type of muffler that is known to give very satisfactory re¬ 
sults when properly proportioned is shown in Fig. 218. It con¬ 
sists of two cast iron heads which are connected by three con- 
centric sheet metal tubes, the two inner ones of which are here 
shown conical, but may, of course, also be cylindrical. The gas 
enters the outer one of the three concentric chambers through a 
spiral passage cast on one of the heads. They expand in this 
chamber and then pass through fine perforations at the opposite 
end in the partition wall into the intermediate chamber, and thence 
through similar perforations at the near end of the innermost 
partition wall into the central chamber, from which they escape 
at the further end. This muffler is a development of an earlier 
















328 STARTING CRANK, MANIFOLDS AND MUFFLER. 


type, also of concentric cylindrical chamber form, in which the 
gases entered the innermost chamber and were discharged from 
the outermost chamber. It has the advantage over the latter 
type that it is easier to provide a comparatively large initial ex¬ 
pansion chamber, and that the walls with which the hot gases 
come in contact are exposed to the atmosphere, hence the gases 
are cooled more rapidly. The swirl given to the gases by the 
spiral inlet passage brings the newly entering charge immediately 
in contact with the entire outer wall of the muffler and prevents 
any direct impact between the column of charge and a wall of 
the muffler. 

Another form of muffler is shown in Fig. 219. This consists 
of a considerable number of flat cylindrical chambers of pressed 
steel with one head integral and the edge at the open end flanged 
out so as to fit over the closed end of the adjacent section. The 
heads are provided with openings through which adjacent cham¬ 
bers communicate, these openings being alternately at the centre 
and near the circumference, thus compelling the gases to follow 
a tortuous course through the muffler. The heads are made of 
cast iron and the gases enter and leave at the centres of the heads. 
The first and last partition walls differ from the others, the first 



Fig. 219.—Baffle Plate Muffler. 


having openings both at the centre and near the circumference 
to facilitate rapid expansion, and the last has fine perforations 
only, so as to break the discharge up into a large number of fine 
streams. 

A third type of muffler, shown in Fig. 220, consists of two ad¬ 
jacent cylindrical chambers which communicate through a large 
number of small tubes secured in a partition wall cast integral 
with a central tube. This tube is cast with a partition wall at the 

centre of its length and with fairly large openings in its wall 

through which the gases enter the first expansion chamber; 

thence they pass through the small tubes already referred to 

into the second expansion chamber, and thence through small 


































STARTING CRANK, MANIFOLDS AND MUFFLER. 329 - 



Fig. 220.—Small Tube Muffler. 


perforations in the wall of the second section of the central 
pipe through which they are discharged. 

In Fig. 221 is shown a muffler of patented design, known 
as the Ejector muffler, which is used on a good many American 
cars. It consists of three expansion chambers formed by conical 
diaphragms, of which alternate ones are perforated near the 
centre and the outer edge, respectively. There is an axial tube 
leading through the muffler, the diameter of which is reduced at 
about the middle of its length. Part of the gas entering the 
muffler passes directly from the axial tube into the central 
chamber and through the second set of cones, while the rest 
enters the first chamber and passes first through the first set 
of cones and then through the second. A small portion of the 
exhaust is carried straight through to a nozzle at the farther 
end of the muffler, through which it is discharged to the at¬ 
mosphere. This discharge, by reason of the familiar injector 
principle, creates a partial vacuum in the third expansion cham¬ 
ber, thus creating a suction which draws the gas out of this 
chamber in a more or less uniform stream. The gas passing 
through the expansion chambers naturally reaches the third 
chamber somewhat later than the gas passing through the axial 
tube reaches the nozzle, and the pressure in the third chamber 
at the moment when the gas from a new explosion enters that 



Fig. 221.—Ejector Muffler. 















































































330 STARTING CRANK, MANIFOLDS AND MUFFLER. 

chamber is therefore always below atmospheric, which insures 
low back pressure. 

Back Pressure of Mufflers—Several competitions and tests 
with the object of improving the design of mufflers have been 
held under the auspices of the Automobile Club of France. 
In a competition held in 1903 the tests were made on a single 
cylinder 8 horse power motor. A two-way valve was inserted 
in the exhaust pipe by means of which the exhaust could be 
turned either into the atmosphere or into one of the mufflers. 
In these tests, of the four mufflers which gave the most satis¬ 
factory results, from the standpoint of silence, three absorbed 
from 11 to 12 per cent, of the power, while one absorbed as 
much as 20 per cent. 

The second of these trials, conducted in 1905, brought out 
the interesting fact that a length of pipe connected to the ex¬ 
haust manifold of the motor may increase the power output. 
During the first part of the trial all tests were made with the 
mufflers connected directly to the exhaust manifolds. It was 
then decided to place them at some distance from the motor, 
to reproduce the conditions on an automobile, and the con¬ 
nection was made by a pipe with three bends in it, 6 feet long 
and of 2 inches internal diameter, substantially the same as 
the manifold diameter. The engineer in charge of the tests 
anticipated that a certain correction would have to be made for 
the loss in this pipe, and proceeded to determine this loss. He 
was greatly surprised to find by repeated experiments that this 
pipe increased the power, and accurate tests showed the in¬ 
crease to be 1.4 per cent. The explanation offered for this 
phenomenon is that when the gases pass through the 6 foot pipe 
their temperature is reduced to quite an extent and they issue 
from the mouth of the pipe at a considerably lower speed 
than if discharged directly from the exhaust port, hence the 
resistance to their entrance into the atmosphere is less, which, 
of course, means less back pressure. The most silent mufflers 
in this competition reduced the power of the motors 10 to 11 
per cent. 

The final outlet from the muffler is generally through a pipe 
of considerably less cross-section than the exhaust pipe, which 
outlet pipe is sometimes tapered or reduced in section away 
from the muffler. Some mufflers arranged crosswise of the frame 
at the rear are provided with two outlet pipes, slightly in¬ 
clined downwardly and with their outlet ends flattened so that 


STARTING CRANK, MANIFOLDS AND MUFFLER. 331 


the exhaust is discharged in a sheet at the rear of the car, 
whereby the dust is kept down to a certain extent. 

The chief requirements in muffler design may be summarized 
as follows: Use non-sonorous materials, such as soft sheet 
iron and cast iron and no brass or aluminum. Provide an ex¬ 
pansion chamber of large size as compared with the volume 
of the individual cylinder. Keep the hot gases during the first 
part of their passage through the muffler in contact with sur¬ 
faces exposed to the atmosphere so as to reduce their tem¬ 
perature and pressure as quickly as possible. Gradually decrease 
the cross sectional area of the passages so that the exhaust 
will issue from the outlet pipe in a practically continuous stream. 


CHAPTER XV. 


THE FLYWHEEL. 

As has already been explained, in a four cycle motor there 
is only one explosion or power stroke in every four strokes 
of the piston. Consequently in a single cylinder motor a 
turning moment is impressed upon the crankshaft during 
only one stroke out of four, and there are three idle strokes 
between power strokes. 

In order to have a practically usable motor the turning 
moment or torque must be substantially uniform throughout 
the cycle, and this makes it necessary to store up some of the 
mechanical energy developed during the power stroke and 
draw upon this store of energy during the three idle strokes,, 
so that the rate of doing external work is practically constant. 
The flywheel forms this means of storing energy. 

Variation of Crank Moment—Fig. 222 shows a crank 
moment diagram for a single cylinder 4x5 inch motor, based 
upon indicator diagram Fig. 76, and a 10 inch connecting rod. 
The diagram shows the crank moment at low speeds, and the 
effect of the inertia of the reciprocating parts on the crank 
moment is neglected, which is the only thing that distin¬ 
guishes this diagram from that of Fig. 19. The inertia of 
the reciprocating parts is here neglected partly for the sake 
of simplicity and partly because the flywheel function is most 
important when the engine runs at low speed, when the inertia 
effects of the reciprocating parts are negligible. 

The cylinder develops approximately 415 foot-pounds of 
energy per explosion, net; that is, the indicated energy of the 
explosion stroke is 474 foot-pounds and that of the compres¬ 
sion stroke 59 foot-pounds, leaving 415 foot-pounds of avail¬ 
able energy. Now consider a crank with an arm 1 foot in 
length. The above amount of energy— 415 foot-pounds—is 
expended during two revolutions of the engine crankshaft, 
and during these two revolutions the outer end of the i-foot 
crank travels 

2x2x3.14=12.56 feet . 


332 



THE FLYWHEEL 


333 



— / dO * -— 1 - /< 30 ° - 1 - / Q 0 ° - 1 - / ao ° — 

Fig. 222.—Low Speed Crank Moment Diagram of Single Cylinder 4x5 Inch Motor. 
































334 THE FLYWHEEL. 

The force acting at the outer end of the i-foot crank arm is 
therefore 


~ 33 * ounds ( a PP r -) 

Consequently, the average turning effort or torque is 33 
pounds-feet. We may consider the resisting moment of the 
load on the motor to be constant, and its value, then, is 33 
pounds-feet. The resisting moment will very nearly be con¬ 
stant when the car is pulling up a steep hill at a very slow 
pace. 

Referring now to Fig. 222, it will be seen that while the re¬ 
sisting moment of the motor is only 33 pounds-feet, the torque 
impressed upon the motor crankshaft during the power stroke 
rises as high as 320 pounds-feet. The excess torque goes to 
accelerate the flywheel and to store up energy in it. In the 
figure, the area ABODE represents the total energy of one 
explosion = 474 foot-pounds. The area BCD represents ex¬ 
cess energy developed during the first four-fifths of the ex¬ 
plosion stroke, which is stored in the flywheel. This area rep¬ 
resents 382 foot-pounds. The energy stored up in the fly¬ 
wheel during the power stroke is given out again during the 
three so-called idle strokes, doing the external work during 
these three strokes, represented by the rectangle E F G H, 
and the work of compression represented by the area H I J. 
In reality the flywheel begins to supply some of the energy 
necessary to perform the outside work at the point D, be¬ 
fore the end of the power stroke. The area D F G H I J E, 
representing the total energy given out by the flywheel during 
one cycle, is exactly equal to the area BCD, representing 
the energy absorbed by the flywheel during one power stroke. 

Constructions similar to Fig. 222 enable us to determine the 
relative amount of flywheel capacity theoretically required for en¬ 
gines with different numbers of cylinders. In Fig. 223 is shown 
a crank moment diagram for a double cylinder opposed 4x5 inch 
motor, other conditions being the same as in Fig. 222. The 
crank moment during only one revolution is shown, as the 
diagram repeats itself every revolution. Measurement shows 
that in this case the flywheel has to absorb 302 foot-pounds 
of energy during the explosion stroke and give it out during 
the compression stroke. Fig. 224 shows a crank moment 
diagram of a four cylinder 4x5 inch motor for one-half revo¬ 
lution, this diagram repeating itself every half revolution. 
It will be seen that in this case the flywheel must store up 



THE FLYWHEEL. 


335 


169 foot-pounds of energy during the first half of every stroke 
and give it out again during the last half. Fig. 225 shows a 
crank moment diagram for a six cylinder 4x5 inch motor for 
two-thirds of a revolution. In this case the flywheel must 
store up and deliver periodically 82 foot-pounds. The fly- 



Fig. 223.—Low Speed Crank Moment Diagram of Double 
Cylinder Equally Timed 4x5 Inch Motor. 

wheel capacities required with engines of different numbers 
of equal sized cylinders therefore compare as follows: 


Single cylinder. 100% 

Double cylinder, equally timed. 80% 

Four cylinder. 44% 

Six cylinder. 22% 
























336 


THE FLYWHEEL. 


Effect on Flexibility and Starting of Motor and Car— 
When these relative flywheel capacities are used the different 

motors will run with the same degree of irregularity, or have 
the same coefficient of speed fluctuation—provided the resistance 
is constant. There is, however, another way of looking at the 



Fig. 224.—Low Speed Crank Fig. 225.—Low Speed Crank 
Moment Diagram of Four Moment Diagram of Six 

Cylinder 4x5 Inch Motor. Cylinder 4x5 Inch Motor. 

matter, namely from the standpoint of vehicle starting. The 
throttle cannot be fully opened before the clutch is let in, as 
otherwise the engine would be speeded up excessively. Con¬ 
sequently if it is desired to get away fairly quickly the fly¬ 
wheel must furnish a good deal of energy. If the flywheels of 






































THE FLYWHEEL. 


337 


four and six cylinder motors were made as small as indicated 
by the above comparative table there would be considerable dan¬ 
ger of stalling the motor when letting the clutch in rather 
quickly, especially if the power of the motor is low as com¬ 
pared with the weight of the vehicle and the low gear reduction 
ratio is not very great. The flywheels of these multicylinder 
motors are therefore generally made larger than indicated by 
the comparison. Moreover, when using four or six cylinders one 
naturally wants a very flexible motor which will run through a 
wide range of speed under full torque. With a single cylinder 
motor this is not well possible, as it would require an enormously 
heavy flywheel, but with a multicylinder motor the end can be 
attained with moderate flywheel capacity, and is, therefore, within 
the limits of practical possibility. 

Another reason for making four and six cylinder flywheels 
somewhat larger than required by running conditions of the 
motor is that it facilitates starting of the motor. Without a 
flywheel it would be impossible for the operator of average 



1 

—1— 
7 

L. 

-J 

L- 




Fig. 226. 

strength to turn the crankshaft of a motor of, say, 5 inch bore, 
against compression. The provision of a heavy flywheel enables 
him, when spinning the motor, to supply the requisite energy 
for compressing the charge at a practically uniform rate during 
two revolutions of the flywheel. The greater the flywheel ca¬ 
pacity the less will be the fluctuation of the cranking resistance 
at a certain average cranking speed. 

Energy Stored in Flywheel—The kinetic energy of a mov¬ 
ing weight, is 

£ — foot-founds 

2 g 

when w is given in pounds, v in feet per second and g is 
the acceleration of gravity = 32.16 feet per second per second. 
Now, referring to the sketch Fig. 226, let w be the weight of a 

small particle of matter in the rim of a flywheel at a distance — 

12 

feet from the centre of rotation, and let the flywheel rotate with 













/}sci •$ of ffotation^ 


338 


THE FLYWHEEL. 


an angular speed of — revolutions per second. 1 hen the veloc- 

6 o 

ity of the particle iv is evidently 

-L — feet per second , 

o T 12 60 

and the energy stored up by the particle w is 


E = 


4 zv x 2 r 2 w 2 


Zt/ 7T Z r w 1 


2^ X I2*X 60 2 259,200^ 


foot-pounds . (75) 


The angular speed of the motor is generally expressed in 
radians per second— 




2 7 X 71 
60 ' 


and substituting this in equation (75) we have 
.2 


1 zv r l 

h — & 


288^ 


( 76 ) 



ixj y * tju l y* \ “ # 

The expression- or -{ — ) is known as the 

g \I2 / 

polar moment of inertia ( 7 p ) and is a measure of the flywheel 
capacity. In calculating this dimension, w is expressed in pounds; 

— represents feet, and g is expressed in feet per second per 
12 

second, hence the polar moment of inertia or flywheel capacity 
is expressed in feet-pound-seconds 2 . The stored energy may 
therefore also be expressed as follows: 



2 


Now consider a flywheel with a rim of rectangular cross 
section, as shown in Fig. 227. Take a thin cylindrical shell of 
radius r inches and thickness dr. All the metal in this thin 
shell may be considered at a uniform distance r from the centre 































/ 


THE FLYWHEEL. 339 


of rotation. If the flywheel is made of cast iron—the common 
material—which weighs 0.26 pound per cubic inch, the weight of 
metal in the shell is 

zv = ztt r adr X o.26 = o. 52 7r a ?' dr founds , 
and the velocity 


* 2 — r — feet per second; 
12 X 60 


hence 


* 2 = 


o o o 

4 7r - r n 
144 X 3600 


and the energy stored up 


7r 2 r 2 n 2 


E - 


O. 52 tt a r dr X L 

_ • 144 X3600 

0.52 7 r 3 w 2 a r 3 dr 


259,200 g 


( 77 ) 


In order to find the energy stored up in the entire rim we must 
integrate this expression between the limits r\, the intemaJ 
radius, and r 0 , the external radius. 


-a r \ H 

f * O.52 ir J 

n 2 a r 7, dr 

J 259200 g 

To 

0.52 7r 3 a 

259.200 jnr ^ 

f 'r z dr 

' To 

0 .52 7T 3 U 2 a ( 

r o r \ \ 

259,200,? \ 

4 / 


2,068 ”«» ('•*« _ r ') f° ot t° unds 


( 18 ) 


In practical work sufficient accuracy is obtained by considering 
the entire weight of the rim concentrated at a distance from 
the axis of rotation equal to the arithmetical mean between the 
inner and outer radii of the rim. The weight of the rim is 


xv 


— (r ? 0 — ?' 2 i^ ~ci X 0,26 pounds, 


and the mean radius is 


^ + n . , 

inches. 

2 

The linear velocity of a point at the middle of the rim is 
therefore 
















340 


THE FLYWHEEL. 



feet per second. 


After these two values have been found they are inserted in 
the formula for kinetic energy and the calculation is carried 
through. As a rule only the energy stored in the rim is calcu¬ 
lated with, though, of course, the spokes or web also have some 
flywheel effect. 

Dependence of Flywheel Capacity Required on Bore and 
Stroke—In the foregoing we investigated the flywheel capacities 
required in motors with different numbers of cylinders. We will 
now consider the problem of how the flywheel capacity should 
vary with the bore and stroke, the number and arrangement of 
the cylinders remaining the same. Take a single cylinder motor, 
for instance. All the energy is developed during the power 
stroke, but the motor does work at a substantially uniform rate 
during all four strokes. The flywheel, therefore, must store up 
about three quarters of the energy liberated by the explosion. In 
reality the fraction is a little higher, because during the compres¬ 
sion stroke the flywheel must supply the energy not only for 
doing the outside work but also for compressing the charge. 
The point is that a certain fixed percentage of the energy of 
each explosion is stored in the flywheel, and the fluctuation in 
its speed caused by the absorption and return of this energy must 
not exceed a certain percentage. Now, the energy of an ex¬ 
plosion is directly proportional to the piston displacement or to 
the expression b 2 l ( b being the bore and l the length of stroke). 
We have already found an expression (76) for the energy storing 
capacity of the flywheel, and we may therefore write 


b 2 l 


zv r 2 

w 2 - 

288^ 


( 79 ) 


But the angular velocity w increases as the length of the 
stroke / decreases, and vice versa. It was formerly considered 
that the two factors varied in direct inverse ratio, but it is 
now generally admitted that a long stroke motor will run at 
higher piston speeds than a short stroke one, and it is therefore 
advisable to write 


CO 




Inserting this value in proportionality (79) we have 

b'l ~ - ^ 

l 288^ 






THE FLYWHEEL. 


341 


6 2 / 2 ~ 


zu r 2 
288 g 


(80 


An analysis of considerable data regarding flywheel dimensions 
on hand shows that the average proportions are as follows: 

For single cylinder motors 

zu r 2 b 2 1 2 

288 g 875 

For doable cylinder motors exploding at equal intervals 

zu r 2 b 2 l 2 
288 g 1000 
and for four and six cylinder motors 

zu r 2 o 2 l 2 
288 g 1100 


These equations can be simplified so as to give the weight in the 
rim directly— 

For single cylinder motors 


zu = 10.6 



pounds 


(81) 


for double cylinder motors with explosions equally spaced 


zu = 


— m 


pounds .. (82) 


and for four and six cylinder motors 


zu == 8.4 


bj' 


pounds ...(83 


The outside diameter of the flywheel is a function of the 
length of stroke and is generally made 

<*0 = 3 - 3 / . (84) 

However, in determining this diameter the road clearance has to 
be taken into consideration, since the flywheel is usually the 
lowest part -on the car, and on American roads a clearance of 
about 10 inches at the middle of the chassis is essential. It is 
to be presumed that now that long stroke motors are coming 
into favor the ratio between flywheel outside diameter and piston 
stroke will be reduced, since it is hardly practical to make the 
outside diameter more than 20 inches on a touring car. The 
advantage of large diameters in flywheels is clearly shown by 
equations (81) to (83), the required wheel rim weight being in¬ 
versely proportional to the square of the mean radius of the 
rim. 

The above equations give the average weight of flywheels as 
















342 


THE FLYWHEEL. 


fitted at present. It should he pointed out that a heavier fly¬ 
wheel makes it easier to spin the motor in starting, more diffi¬ 
cult to stall the motor by letting in the clutch a little too rapidly, 
makes the motor more flexible and gives a more uniform torque, 
thus reducing the stresses in the transmission members and on 
the tires. Therefore, wherever a little extra weight is not of 
great importance, it is advisable to make the flywheel somewhat 
heavier than the weight given by the above formulae. Also in 
motors with unusually high compression a slightly greater fly¬ 
wheel capacity than found from the equations should Be pro¬ 
vided, because the energy of an explosion increases with the 
compression pressure, and so does the work of compression. 

A four cylinder 4x5 inch engine should have a flywheel with 
a mean radius of gyration of about 7.5 inches and the weight 
in the rim required according to equation (83) would be 

g ^ ( 4 X _5 \ '=60 pounds. 

\ 7*5 / 

Now, suppose that the resistance of the load is substantially 
constant, as it would be in pulling slowly up a steep hill on the 
low gear, requiring the entire torque of the motor. At what 
speed would the motor become stalled? 

We found previously that in a motor of this size when running 
under full throttle the flywheel alternately stores and gives out 
169 foot-pounds. The motor will stall, of course, when the maxi¬ 
mum speed during a cycle, which is attained just before the end 
of the power stroke, is such that the energy stored in the fly- . 
wheel is 169 foot-pounds. We then have 

/ \ 2 zu r 2 

\ 2ir, 7 = 169 

from which we find n = 3.5 revolutions per second or 210 revolu 
tions per minute. Since the motor speed varies between this, 
and zero during a cycle, the average speed just before stalling 
will be approximately half this or 105 revolutions per minute. 
Of course, the motor will run very irregularly at * speeds only 
slightly exceeding this figure. 

Centrifugal Force on Flywheel—When a flywheel revolves 
about its axis every particle of matter in it is subjected to cen¬ 
trifugal force which acts radially outward. The value of this 
force has already been given as 

F = 1.226 zv rr r, 

where « is the number of revolutions per second and r the radius 
of rotation in feet, It will, however, be better to continue to 
express n in revolutions per minute and r in inches, in which 



THE FLYWHEEL. 


343 


case we have 


.. u> n 2 r 

/ c = - 

35.240 

1 he weight 70 of the' rim is approximately 

2 7 T r h a X o. 26 pounds. 

where /i is the radial depth of the rim in inches. Inserting this 
value in the equation for the centrifugal force we have 


F c = 


7 r r ' 1 71 2 /e a 


(appr.y 


68 000 

1 his centrifugal force tends to rupture the rim in a plane 
passing through its axis. The component of the centrifugal force 
perpendicular to this plane is equal to one-half the total force 
multiplied by the ratio of the diameter to half the circumfer¬ 
ence 


r. _ tt ?- 2 71 2 ha r 2 r r 2 w 2 h a 

r\\ -X ^ X -• —- 

68,000 2 7r r 68,000 

1 his tension is taken up by two cross sections at opposite 
ends of a diameter whose combined area is 2 ha square inches. 
The unit stress therefore is 


r 2 71 2 h a 

c 68,000 r 2 7 i 2 , , . , 

o = ■■■ < —- pou7icxs per square nick. 

2 h a 136.000 

Taking the tensile strength of ordinary cast iron to be 16,000 
pounds per square inch, a flywheel would burst when running 
at a speed 

71= t /16,000 X 136,000 , , . 

n y —-- - 2 - revoluhoyis per mi7iute. 

The maximum value of r for automobile flywheels is about 9 
inches, hence the bursting speed is 

/ 16,000 X 136,000 ~ 

a/ — -———- = s. 180 rev. p. tti. 

> 9 X 9 

Of course, if there- are blowholes or similar defects in the 
casting it may burst at a lower speed. The ordinary 4x5 inch 
engine would not be able to exceed one-half this speed, and 
since the centrifugal force varies as the square of the number 
of revolutions per minute, it would then still have a factor of 
safety of 4. A .few manufacturers have used steel flywheels, 
and steel founders have made efforts to introduce these wheels, 
but they stand little chance of general adoption, owing to 
their high cost. In regard to the safety of flywheels, it should be 
pointed out that this is further increased by a solid web, as very 1 
generally used. Where steel flywheels are used the outside diam- 












344 


THE FLYWHEEL. 


eter is generally made somewhat greater than in the case of cast 
iron wheels. Thus the 110x140 mm. (44x5.6 inch) Mercedes 
motor has a cast steel flywheel with an outside diameter of 
about 21.5 inches. 

Fluctuations in Car Speed—In connection with the flywheel 
capacity of the motor the flywheel effect of the car itself when 
going at speed must be taken into account. It was assumed so 
far that the resisting moment on the motor was constant This, 
however, is by no means the case when the car is traveling at a 
fairly high speed. It then has stored up a great deal of kinetic 
energy, upon which it draws immediately the motor slows down 
in speed. A couple of examples will show the fluctuations in ve¬ 
hicle speed under more or less extreme conditions. 

Referring to Fig. 21, which represents the crank moment dia¬ 
gram of a four cylinder 4x5 inch motor at normal speed, we find 
by measurement that the excess energy given out by the motor 
during part of each stroke is 140 foot-pounds. Owing to the in¬ 
ertia of the reciprocating parts the motor gives out this excess 
energy during the last half of each stroke, instead of during 
the first half, as might be expected. Consequently the motor and 
the car will actually slow down immediately after an explosion 
and speed up just before an explosion. The 140 foot-pounds 
excess energy will speed up both the flywheel and the whole car. 
Since the flywheel is rigidly connected to the driving wheels the 
speeds of the two will increase and decrease in the same pro¬ 
portion. 

Denoting the weight and velocity of the car by capitals and the 
weight and velocity of the flywheel rim by small letters, we have 

Total energy stored up at normal speed: 


W V 2 -f zv v 2 

_|_• 

2 g 


Total energy stored up at minimum speed: 


W Vf + tu 
2^ 


We know that 

(WV 2 -\-zvv 2 IV Vf-\-zvvf 


( 



= 140 foot-founds; .(85) 


\ 2 g 

also that 


2 g 



from which it follows that 



Substituting in equation (85) 







THE FLYWHEEL. 


345- 


W V 2 + tv v 2 — W V x 2 — w 


VI 

V 2 


v • 


2 ^ 


= I4O. 


( 86 ) 


We will assume that the weight W of the car with passengers 
is 2,500 pounds and the weight w of the flywheel rim 60 pounds. 
The flywheel rim has a mean diameter of 15 inches and its linear 
speed when the motor turns at 1,200 revolutions per minute is 

3.1416 X 15 X 1200 

I2 6o = 7 8 • 5 f eet P er second. 

We will suppose that the car is provided with 32 inch driving 
wheels, and that these are geared in the ratio of 3:1 on the high 
gear. The car speed then is 

1200 32 

3 X _ 6o ^ 12 ^ 3 ' 14 = 55 ' 8 feet per second. 

Inserting the values thus found in equation (86) we have 

2500X 55 • 8 2 + 60X 78 • 5 2 — 2500 X 2 — 60 78.5 ? = 2X32 • 3 X14° 

55 *8 2 

which when solved gives 

V x = 55.77 feet per second. 


Similarly we find that the maximum speed, near the end of the 
explosion stroke, is 55.83 feet per second. Consequently the car 
speed ranges between 55.77 and 55.83 feet per second, a variation 
of less than one-tenth of 1 per cent. 

Now consider that the car is pulling up a steep hill on the 
low gear, which may be considered to be one-fourth as high as 
the high gear. Also let the engine be pulled down to 400 revo¬ 
lutions per minute. The car speed V is then 

55-8 

——— = 4.65 feet fer second 
3X4 » 

and the flywheel rim speed v 


78.5 


= 26.17 feet per seco ? id . 


Substituting these values in equation (86) we have 

v \ 

2500 X4-65 2 + 6oX 26.17 2 — 2500 X 2 —60^^226. i7 2 = 169 X 7 . g t 

which when solved gives 

^1 = 4.38 feet per second. 

In the same way we find that the maximum momentary speed 

V 2 = 491• 

The speed fluctuation under these conditions is, therefore, 

(4.91 — 4.38) 100 


4-65 


= 11 . A^fer cent . 












346 


THE FLYWHEEL. 


of the average. The fluctuation would be still greater if 
under the above running conditions one of the cylinders should 
miss an explosion. It can be easily shown that in that case 584 
foot-pounds of energy would be withdrawn from the system and 
the speed would drop to 3.62 feet per second. A sudden drop in 
the speed of the car from 4.91 to 3.62 feet per second would, of 
course, be very distinctly and disagreeably noticeable. 

Rim and Web or Spokes—The flywheel generally serves 
as one member of the friction clutch, and this determines its 
shape to quite an extent. If a conical clutch is used the rim of 
the flywheel is generally bored out on the inside to form a conical 
surface whose generatrices make an angle of about 7.5 degrees 
with the axis of the cone. Fig. 228 shows a sectional view of a 
typical flywheel chambered out for a conical clutch. If a multiple 
disc clutch is used a housing for the clutch is generally cast inte¬ 
gral with the flywheel, of a considerably smaller diameter than 
the rim. In proportioning the rim it is advantageous to make it 
rather wide faced and not very deep radially, so as to make the 
mean diameter as nearly equal to the outer diameter as possible. 
In American practice most flywheels are now made with solid 
webs of half an inch thickness. If the flywheel is bolted to an 
integral flange on the crankshaft the web is generally formed with 
a circular offset which is bored out to receive the flange, so as to 
accurately centre the flywheel. The number and size of bolts 
to use and their distance from the axis was discussed in the chap¬ 
ter on the Crankshaft. It may here be repeated that it is advan¬ 
tageous to use dowel bushings around the bolts. 

Keying Flywheels—Occasionally the flywheel is secured to 
the crankshaft by means of a key. On the whole this method is 
not very satisfactory, as the key is apt to loosen sooner or later 
and to cause a very disagreeable knock. Where this construc¬ 
tion is used the key should be made as large as it can con¬ 
veniently be made. In some constructions of this type the shear¬ 
ing area of the key is 

tv r 2 
a =- 

1350 

where w is the weight of the rim in pounds, r the mean radius of 
the rim in inches, and d the diameter of the shaft where the fly¬ 
wheel is keyed to it. In stationary gas engine practice, where 
flywheels are commonly keyed to the crankshaft, the hubs are 
made of a diameter equal to twice the diameter of the shaft 
and of a length equal to 1.75—two times the diameter of the shaft. 
The width of thp key is made about one-fifth the shaft diameter 


THE FLYWHEEL. 


34? 


Since the materials are the same, these proportions should also 
be best for automobile practice, and the few instances of keyed 
flywheels of which we know accord very well with these rules. 

One plan resorted to in order to increase the strength of keyed 
joints of cast iron flywheels is to provide them with a steel 
flanged hub to which the web of the flywheel is bolted in the 
usual way, as shown in Fig. 228. Steel is superior to cast iron in 
respect to both compressive and shearing strength, and with a 
certain size of key the joint will un¬ 
doubtedly be more secure if it is between 
steel and steel than when it is between 
steel and cast iron. It is customary when 
keying a flywheel to the crankshaft to 
taper the bore of the hub and the seat 
for it on the shaft one-eighth of an inch 
per inch in length, and drawing the fly¬ 
wheel up on its seat by means of a nut. 

Fan Wheels—A few manufacturers, 
following a practice first introduced by the 
Daimler Motor Company, of Germany, cast 
their flywheels with fan shaped spokes, so 
that it will aid the radiator fan in creating 
a draught through the radiator (see P'ig. 

229). Some makers place the flywheel at 
the front end of the motor directly back 
of the radiator and depend entirely upon 
its fan shaped spokes for producing the 
radiator draught. Fan flywheels are used 
in conjunction with both cone and mul¬ 
tiple disc or band clutches. The cone clutch 
is generally about the same diameter as the 
flywheel rim, and the clutch cone must then 
be provided with spokes, so the air moved 
by the flywheel may pass through it. Oc¬ 
casionally the clutch spokes are also made 
fan shaped, so as to assist the flywheel 
spokes rather than act as an obstruction to the air moved by 
them. 

If the multiple disc clutch is used, this is generally only of 
about half the outside diameter of the flywheel rim. The fly¬ 
wheel is then cast with the clutch housing in the centre and fan 
spokes connecting the outside of the housing with the inside of 
the rim. The spokes are preferably laid out helix fashion. 
Consider a median surface through the spoke oarallel with iU 



Iron Flywheel 
W ith Steel 
Hub. 


























348 


THE FLYWHEEL. 


sides. The line of intersection of this surface with the inner 
surface of the rim cuts the generatrices of the latter surface 
at a constant angle. There does not seem to be much uniformity 
in the design of these spokes, the “angular advance” or “lead” 
of the helix varying between wide limits. For average sized 
engines a lead of 40 inches seems to give good results. That 
is, if the flywheel rim were extended to a great length and the 
spoke the same, the joint of the spoke with the rim would make 
a complete revolution in 40 inches measured along the axis of 
the flywheel. Each point of the intersection of the helix with 
the inner surface of the rim connects to the axis by a straight 



Fig. 229.—Fan Flywheel. 

line perpendicular to the axis. These flywheels are generally 
made with either five or six spokes. 

Machining Flywheels—Flywheels can be machined to advan¬ 
tage in a vertical turret lathe. A method of machining a typical 
form in a Bullard lathe is illustrated in Fig. 230. The flywheel 
is mounted on the revolving table, first with one face up and then 
with the other, and with these two settings the wheel is bored 
out and machined all over. A cast iron flywheel of the type 





















THE FLYWHEEL. 


349 



Fig. 230.—Machining a Flywheel. 






































































































































































350 


THE FLYWHEEL. 


shown, about 18 inches in diameter, can be completely machined 
in about thirty minutes. The illustration clearly shows each of 
the successive operations. 

Balancing Flywheels —The flywheel is by far the most im¬ 
portant rotating part of the motor, and in order that the latter 
may run smoothly and without’’ undue vibration, the former 
must be carefully balanced. It is not sufficient that it be in 
static balance; that is, that it will remain in any position when 
mounted on a shaft and placed on two balance ways; it 
must also be in running balance. Of course, .if it is machined 



Fig. 231.—Defiance Flywheel Balancing Machine. 

all over, and the work is carefully done, it is almost certain 
to be in true running balance. Not so if the wheel is spoked, 
for instance, and cannot be finished on and between the spokes. 
In that case it is necessary to test for running balance, for which 
purpose the machine shown in Fig. 231 was designed by the 
Defiance Machine Works, of Defiance, Ohio. It consists of a 
conical frame in the centre of which is placed a vertical ground 












THE FLYWHEEL. 


351 


steel spindle standing in a step bearing in the base of the frame 
and passing up through a long bearing at the top. This spindle 
at its upper end carries a face plate provided with two driving 
pins which project upward parallel with the axis of the spindle. 
A steel centre with tapered shank is fitted into the vertical 
spindle, the upper end of the centre being reduced to a point upon 
which the flywheel to be balanced rests. 

The vertical spindle is rotated by friction gearing from a 
horizontal shaft carried in suitable bearings in the base of the 
machine. The horizontal shaft is in two parts, which are con¬ 
nected by a friction clutch and moved longitudinally by means 
of a counterweighted lever, so arranged that the weight serves 
to press the clutches together when it is desired to rotate the 
vertical spindle, and to disconnect the clutches when it is de¬ 
sired to examine, test and mark the flywheel while running alone,, 
as the driving pins are liable, when pressing against the fly¬ 
wheel, to produce uneven rotation. The true unbalanced condi¬ 
tion is best shown when the disturbing influence of the motive 
power is withdrawn. 

The method of using this machine is as follows: First a stand¬ 
ing balance is produced by applying a weight to the inside of the 
rim. Then the flywheel is rotated and the edge of the rim 
marked with a pencil of moistened chalk. If the mark occurs- 
within one-fourth of the circumference of the wheel from the 
weight in the direction of rotation the weight is raised and at the 
same time advanced toward the mark. If the mark occurs at 
more than one-fourth the circumference from the weight the 
weight is lowered and placed further back from the mark. If the 
edge of the wheel is reached in this way and the wheel still runs 
out a counterweight should be placed diametrically and trans¬ 
versely opposite. The weight should always be so adjusted that 
the standing balance is not disturbed. Where it is not permissible 
to attach weights material must be removed from the rim in 
order to effect a balance, and the reverse process must then be 
followed, but frequently the use of trial weights facilitates the 
solution. 


CHAPTER XVI. 


SPEED CONTROL—THE GOVERNOR. 

The speed of an automobile motor is controlled by means of 
a throttle valve in the gas inlet pipe, which is generally of the 
butterfly type, but occasionally—especially on foreign cars—of 
the barrel type. As a rule the throttle forms an integral part of 
the carburetor, and its design, therefore, need not be discussed 
here. The throttle valve is connected to a finger lever located on 
top of the steering wheel or underneath it on the steering column, 
and frequently also to an accelerator pedal, so that the driver 
can vary the gas supply to the motor at will, and hence control 
its speed. One method of arranging the connections between the 
throttle arm and the throttle lever and accelerator pedal respec¬ 
tively, is as follows: To the throttle arm is connected a spring 
which always tends to close the throttle. The throttle lever con¬ 
nects to the throttle arm by means of a rod having a sliding 
joint, so that the throttle can be opened beyond the position for 
which the throttle lever is set without disturbing the setting of 
that lever. The accelerator pedal is held in the closed position 
by means of a spring and connects to the throttle arm by means 
of a rod in such a way that the arm can move through a certain 
range without moving the rod. This is accomplished by passing 
the rod through an opening in the throttle arm and providing 
it with stops on both sides of the throttle arm but at some dis¬ 
tance therefrom, and interposing a spring between the throttle 
arm and one of the stops. 

Whenever the load on the motor is decreased the motor has 
a tendency to speed up. This tendency to “race” is especially 
strong when the clutch is withdrawn and, therefore, all load re¬ 
moved. In order to prevent the motor from unduly speeding up 
under these conditions the throttle must always be closed im¬ 
mediately. It is a well known fact that careless drivers will dis¬ 
regard this rule, thereby injuring the motor. For this reason the 
motors on those types of cars which, as a rule, are not driven 


352 



SPEED CONTROL—THE GOVERNOR. 


353 


by the owners themselves, are frequently fitted with governors 
which keep the speed of the motor within certain limits. At 
present the great majority of truck motors, and also the motors 
on some of the highest grade pleasure cars, are fitted with 
governors. 

There are two general types of governors in use on automobile 
motors, viz., the centrifugal type and the hydraulic type. We will 
discuss these two types in succession. Whichever type is used, 
it is connected to the throttle valve in such a manner that as 
soon as the speed for which it is set is exceeded, it closes the 
throttle valve more or less, so as to keep the speed within the 
predetermined limit. In addition to the automatic control of the 
engine, some form of hand or foqt control is sometimes provided. 
The governor works against a spring, and the hand or foot 
control mechanism is so arranged that it varies the position of 
the spring rest, in which case the governor will hold the engine 
at any speed the operator may desire. 

Centrifugal Governors —The centrifugal governor, in the 
form in which it is most frequently used, consists of two 
weights carried upon one of the revolving parts of the motor, 
generally the camshaft. The weights are carried by bell cranks 
which are pivoted on brackets rotating with the shaft and gen¬ 
erally either formed integral with or fixed to the camshaft gear. 
The free arms of the bell cranks engage with a grooved sliding 
sleeve on the camshaft. This sliding sleeve is provided with a 
second groove into which engages the end of a lever or bell 
crank whose other end connects to the arm of the throttle valve. 
The revolving' weights are being drawn or forced together by 
means of a spring or springs which may either extend between 
them or surround the shaft and exert their pressure against a 
stop collar and the sliding sleeve respectively. Fig. 232 shows a 
centrifugal governor in diagram. We found that the centrifugal 
force on rotating bodies is 


„ _ zu « 2 r 
/ y c — -• * 

35,240 


In the case of the governor this force is balanced by the spring 
force in any position of equilibrium. The law of the coiled ex¬ 
tension spring is that its pull is proportional to the extension. 
Let a be the length of the spring with its connections in the 
free state. Then 2 r — a is the extension, and the spring force is 

F s =c (2 r — a). 


Hence 


zv w 2 r 


= c{2 r — a) 


35,240 




354 


SPEED CONTROL—THE GOVERNOR. 


35,240c a = (70,480c — zv n 2 ) r 

r = 35,^40^ . . ( g 7 ) 

70,480 c — run 

In the above w is taken as the weight of one of the governor 
weights, since the spring tension is equal to the centrifugal force 
on one of the weights only. 

It will at once be seen that r increases with the speed n. The 
speed at which the governor acts is, therefore, necessarily 
higher for small throttle openings or light engine loads than 
for large throttle openings or heavy loads, but the difference 
between the two speeds need not be very great. 



Fig. 232.—Diagram of Centrifugal Governor. 


Now let r\ and r 2 be the smallest and greatest radii of rota¬ 
tion, respectively, of the governor weights and tit and n 2 the cor¬ 
responding speeds of balance. Then, from equation (87), 

70,480 r x c — zv n x r x = 70,480 r 2 c — zv « 2 * r 2 


70,480 ( r 2 — r x ) c = zv (« 2 2 r 2 — n x * r x ) 

(« 2 2 r 2 — w , 2 r x \ . 

c --—7- pounds fer inch . 

70,480 ( r 2 — r x ) 


( 88 ) 


r \ = 


35,240 c a 
70,480 c — zv n x 


70,480 cr x — zv n x 2 r x = 35,240 c a 


also 




























SPEED CONTROL—THE GOVERNOR. 


355 


a __ 7°>4 8og r \ — zun x 2 r l 
35,240 c 


2 r x 


2 r v — 7V V r i 

35.240 c 


35,240 zv (nj r 2 — n x 2 r,) 
70,480 (r 2 — r x ) 

2 n x 2 r 2 (r 2 —r,) 
w 2 2 ^2 — n \ r \ 


(89 


In the above, c represents the tension or force required to ex¬ 
tend the spring 1 inch in length and a the length of the spring 
in the free state, inclusive of the hooks or connections to the 
centres of the weights. Now let it be desired that the governor 
act at 1,000 r. p. m. of the motor or 500 r. p. m. of the cam¬ 
shaft when the throttle is fully open, and at 1,200 r. p. m. of 
the motor or 600 r. p. m. of the camshaft when the throttle 
is near the minimum opening position. The limiting radii of 
rotation of the governor weights may be assumed to be 1.5 and 
2 inches respectively, and the weight w of each governor ball 1 
pound. Substituting in equation (88) 


c — 1 (6°° X 2 - 500 X T »5 )— g g pojinds fcr inch , 

70,480(2 — 1.5) 

and in equation (89) — 

a — ~ s , r ~— 2 X 500 J X 2 (2 —1.5) 

a — 2 a 1.5 —- 2 - 2 w - = 1.56 inches. 

6oo z X 2 — 50o z X 1. 5 J 

The combined length of the connections from the ends of the 
spring proper to the centre of the weights can be kept within 
0.31 inch, in which case the length of the spring in its free state 
should be 1% inches. If there are to be two springs, one at 
either side of the weights, then the force required to extend 
either of them a distance of 1 inch should be 


= 4.9 pounds. 
2 


The problem then consists in designing a helical spring which 
in its free state measures i^4 inches in length, which will extend 
at the rate of 1 inch per 4.9 pounds applied to its ends and which 
will not be overstrained by extending it to 

4 — 0.31 = 3.69 inches. 

This can be done by means of the rules for the design of 
coiled springs which were given in the chapter on Valves and 
Valve Gears. If the springs turn out to be of unsuitable size, 
the assumptions regarding the range of motion of the governor 










356 


SPEED CONTROL—THE GOVERNOR. 


weights or the speed range through which the governor acts 
can be varied and the calculation made over again. 

Hunting—Suppose the governor weights to occupy a certain 
position intermediate between the two limiting positions, and to 
be held in equilibrium by the centifugal force acting outwardly 
and the tension of the springs acting inwardly. Now, let the 
speed of the engine be gradually increased. This, of course, 
will increase the centrifugal force, and there is then a tendency 
for the weights to fly out from the axis of rotation. However, 
there will be no actual outward motion until the excess of cen¬ 
trifugal force over the spring force is sufficient to overcome the 
static friction of the throttle and of the linkage between it and 
the governor weights. Hence the motor will gain some speed 
before the governor begins to act. The moment the speed be¬ 
comes great enough for the centrifugal force to overcome both 
the spring tension and the frictional resistance, the governor 
weights fly out and the throttle begins to close. Now, as soon 
as the throttle and its linkage begin to move, their resistance 
to motion is somewhat reduced, according to the well known 
law that the friction of rest is greater than the friction of 
motion. The result is that the throttle will be closed further 
than it really should in order to keep the speed of the motor con¬ 
stant. The motor will thereby be throttled down in speed, and 
the throttle in turn will be opened wide by the governor. In 
this way the throttle may be opened and closed by the governor 
and the motor in consequence speed up and slow down periodi¬ 
cally, a phenomenon known as “hunting.” Hunting is particu¬ 
larly likely to occur when starting the motor, when its joints 
are stiff. It can be reduced to a minimum by careful fitting so 
that the friction at the bearings of the throttle shaft and at the 
joints in the linkage is very small, and by carefully balancing 
the throttle valve. 

Types of Centrifugal Governors—The typical form of shaft 
governor as used on automobile motors is illustrated in Fig. 233. 
Two bell cranks A A carrying the governor weights BB are 
pivoted to brackets extending from the web of the cam gear. 
The governor weights are connected by helical springs CC se¬ 
cured to them by hooks at opposite ends. The inwardly extend¬ 
ing arms of the bell cranks engage into a groove on the sliding 
sleeve D on the camshaft E. Into a second groove on this sliding 
sleeve engages the end of a lever arm F whose shaft extends 
through and has a bearing in the wall of the cam gear housing. 
This shaft at its end outside the housing carries a long lever G 


SPEED CONTROL—THE GOVERNOR. 


357 


from which connection is made to the throttle arm. The lengths 
of lever G and the throttle arm are so proportioned that the range 
of throttle motion corresponds to the range of the governor. In 



55 

a 

> 

o 

O 


j 

< 

o 

D 

U« 

i—i 
« 
H 


2 : 

w 


U 




ci 


Lin 


the illustration provisions for adjusting the tension of the gov¬ 
ernor springs CC are shown, in the shape of hook screws HH, 
which screw through the projecting ends of the governor weights 














































































































358 


SPEED CONTROL—THE GOVERNOR. 


BB and are locked in place by means of check nuts. However, 
such adjusting means are hardly required and are seldom pro¬ 
vided. This governor is enclosed with the cam gears and re¬ 
quires no special housing. 

Fig. 234 shows a vertical shaft governor as fitted to the Adler 
truck motor. The shaft may be extended at the top and bottom 
to drive such accessories as the timer and the oil pump, if de¬ 
sired. The governor comprises two pivoted weights AA devel- 



Fig. 234.—Vertical Type Governor. 


oped in the form of a bell crank, whose operating arms engage 
with a grooved sleeve B. The counter spring C is located be¬ 
tween this sleeve and the shaft. The sliding sleeve B is also in 
working connection with the lever D. The shaft of this lever 
extends through the wall of the housing and outside the housing 
carries a lever arm E from which connection is made to the 
throttle valve. The governor shaft receives its motion from the 
camshaft through a pair of helical gears, the driven one of 


















































































SPEED CONTROL—THE GOVERNOR. 359 

which is shown at F. Lubrication of the governor is provided 
for by a direct oil feed to the top bearing. 

The governor shown in Fig. 235 was used by the Cadillac 
Motor Car Company on touring car motors some years ago, 
and is an adaptation from marine steam engine practice. The 
centrifugal member consists of a metal ring A which is pivoted 
to the double armed bracket B secured to the vertical shaft C 




Fig. 235.— Ring Type Governor. 


which is driven through bevel gears from the camshaft. When 
the engine is at rest, the ring A is held in an inclined position 
relative to shaft C by means of the spring D extending between 
the two sliding sieves E and F on the shaft C. The sliding 
sleeve F is connected to the ring A through a link G. The 
sliding sleeve F connects through the bell crank H to the throttle 
valve, and the sliding sleeve E connects through the bell crank I 
with the throttle lever on the steering wheel. The anchorage 




















































































































































































360 


SPEED CONTROL—THE GOVERNOR. 


of the governor spring D can thus be moved by the operator 
at will. The sleeve F is moved along shaft C by revolving ring 
A according to the speed of shaft C, thus opening or closing 
the throttle. This forms, therefore, a combined manual and 
automatic engine control. 

Hydraulic Governors—A number of manufacturers, includ¬ 
ing Panhard & Levassor in France and the Packard Motor Car 
Company in this country, have used so-called hydraulic governors 
for governing the speed of their motors. In these the pressure 
of the pump circulated cooling water is made use of to close 
the throttle when the speed exceeds certain limits. A sketch of 
such a governor is shown in Fig. 236. The device consists 



Fig. 236. —Hydraulic Governor. 


merely of a diaphragm chamber A in which there is a diaphragm 
B acting against the plunger C from which connection is made 
to the throttle arm. The cooling water as it comes from the 
pump may be led through the diaphragm chamber, or this cham¬ 
ber may simply be connected by a pipe to the cooling system at 
a point near the pump outlet. The pressure of the circulating 
water increases with the speed of the pump, and hence, with the 
speed of the motor, the relation between the pressure and pump 
speed depending upon the type of pump employed. The throttle 
can be held open by means of a spring. 
































CHAPTER XVII. 


POWER OUTPUT AND OTHER CHARACTERISTICS. 

Many efforts have been made to develop a simple formula, 
giving the maximum power output of gasoline engines, but 
these attempts have been more or less futile, owing to the 
very large number of factors which affect this output. The 
elements of design which determine the maximum output may be 
enumerated as follows: 

Number of cylinders. 

Bore. 

Stroke. 

Compression ratio. 

Form of compression chamber. 

Valve timing. 

Valve port opening. 

Cross section, length and relative straightness of gas pas¬ 
sages. 

Cooling facilities. 

Weight of reciprocating parts. 

Location of ignition points. 

Besides the above, the fit and alignment of the hearings,, 
their relative diameters and lengths, the bearing metals and 
the amount and grade of lubricant used upon them affect the 
output at the crankshaft. It would evidently be impossible to 
embody all of these factors in a power output formula, and 
for this reason all except the number of cylinders and the bore 
and stroke are generally neglected. 

Horse Power Formulae —If we call the mean gas pressure 
per square inch of piston head area during the power stroke, 
reduced by the mean pressure during the compression stroke,. 
p, and if the cylinder has a bore b and the length of stroke is 
l, then the work done in one cylinder during one cycle is 
evidently 

——A/ foot founds. 

4 X 12 


361 




362 POWER OUTPUT AND OTHER CHARACTERISTICS. 

If the motor runs at n revolutions per minute, then there 

will be — explosions in each cylinder per minute, and the 
2 

power developed by each cylinder will be 


7 r nb 2 f l 


H.P. 


' 2 X 4 X 12 X 33,000 

Also, if there are N cylinders, then the power of the motor 
will be 

//. p - * Nnh * ft 
3,168,000 

This is the indicated power—the power conveyed by the ex¬ 
panding gases to the moving pistons. Of this power a cer¬ 
tain percentage is lost in piston and bearing friction, valve 
operation, windage, etc. Representing the mechanical effi¬ 
ciency of the motor by V, we have for the brake horse power 

B.H.P. = Nn * b% " * 

3,168,000 

The linear speed of the piston is 

5 — 2 n l — -^-L ftet per minute. 

12 6 

Substituting this value in the above expression for the brake 
horse power and incorporating the value of ^ in the constant we 
have 


B. H P = 


Ns b 2 v f 

168,000 


(50 


Of these factors N and b are known in advance, but s, the 
piston speed at which the motor develops its maximum power, 
and v p, the mean effective pressure, are not known. If we 
desire a formula for the horse power involving only the bore 
and stroke, then the question arises whether the piston speed 
of maximum output and the mean effective pressure vary 
with the bore and stroke, and if so, in what proportion. 

European Data and Deductions—A very thorough investiga¬ 
tion of these questions has been made by a committee of 
the Institution of Automobile Engineers (Great Britain). As 
regards the mean effective pressure, it would be expected that 
this would increase with the bore, for the reason that in a 
large bore motor there is less cooling surface to the combus¬ 
tion chamber per unit of charge contained, hence less heat 
will be lost through the cylinder walls. This view is un¬ 
doubtedly correct in so far as motors designed with the same 
compression ratio are concerned, and some experimental data 
is cited by the committee to support this point. On the other 










POWER OUTPUT AND OTHER CHARACTERISTICS. 363 


hand, the very fact that there is a relatively larger wall surface 
in a small bore motor, and hence more rapid cooling of the 
charge, permits of using a higher compression ratio. A higher 
compression ratio would make up for the lesser area of cool¬ 
ing surface, so that the mean effective pressure would be sub¬ 
stantially the same in all sizes of cylinders. This is borne out 
by a diagram in which the mean effective pressures of eighty- 
eight motors are plotted against the cylinder bores, which ac¬ 
companied the report of the committee. This diagram showed 
absolutely no variation of the mean effective pressure with 
the bore. Of 124 motors regarding which data were collected 
by the committee, 5.6 per cent, showed a mean effective pres¬ 
sure above 90 pounds per square inch; 15.3 per cent, between 
80 and 90 pounds, 28.2 per cent, between 70 and 80 pounds, and 
21.2 per cent, between 65 and 70 pounds. About 50 per cent, 
of the engines thus showed mean pressures over 70 pounds 
per square inch. The highest pressure found was 9514 pounds. 

The committee, however, was guided by other considera¬ 
tions than these data. Experiments by one English engine 
manufacturer (White & Poppe) had shown an increase in the 
mean effective pressure from 80.6 pounds for a bore of 3.15 
inches to 102 pounds for a bore of 5 inches, and the experience 
of individual committee members was that the mean effective 
pressure increased with the bore “under similar conditions of 
compression.” It therefore decided to accept 68.5 pounds per 
square inch for v p when the bore is 2*4 'inches and 99.5 
pounds when b = 5 inches. The value of the mean pressure 
for any other bore may then be found from the equation 


v p — 130 ('-if-) 


pounds per square inch 


(9i 


The piston speed corresponding to the maximum output 
was determined for 101 engines, and it was found that it 
varied with the ratio r of stroke to bore, as shown in the 
following table: 


Piston Speed at Max. 

No. of Tests. r. H. P, (Ft. P. M.). 

15 . 1.00—1.08 1303 

30 . 1.10—1.20 1240 

24 . 1.21—1.30 1385 

25 . x.33—1.44 14U 

7 . 1.50—1.61 1595 


The committee found that the piston speed corresponding 









364 POWER OUTPUT AND OTHER CHARACTERISTICS. 


to the maximum horse power may be represented by the 
equation: 


s = 600 (r -f- 1) feet fer minute .(92) 

This gives a piston speed of 1,200 feet per minute when r- I, 
and 1,800 feet per minute when r = 2. 

If now we substitute the values of v p and s just found in 
equation (90) we get 

B . H. P. = 0.464 Nb (b — 1.18) (/ + b ) .(93> 

Substituting for r its value ~ we have 


B. P. P, = 0.464 ( b — 1.18) (/+ b) 

The committee, however, recommends to use the coefficient 
0.45 instead of 0.464, for the sake of simplicity, and its pro¬ 
posed formula is 

B. H. P. = 0.45 (b — 1.18) (/ -f b) .( 94 > 

This formula is intended to give the horse power which can 
safely be obtained from an engine of given cylinder dimen¬ 
sions under the most favorable conditions, and Dugald Clerk 
in a note on the report of the committee said that, in order 
to obtain the power usually developed by a motor on the road, 
the results of the formula should be multiplied by 0.6. 

A. L. A. M. Rating Formula—In this country and England 
a rating formula known here as the A. L. A. M. and abroad 
as the Royal Automobile Club formula, has long been used. 
It is as follows: 


B . 


H. 


P . 


Nb 2 

2-5 


(95 > 


at 1,000 feet piston speed per minute. 

When this formula was first adopted, about 1905, most up- 
to-date motors gave their maximum output at about 1,000 
feet piston speed per minute. In actual road work this piston 
speed is even now seldom exceeded, except in engines built 
more particularly for racing, or of exceptionally long stroke. 
From equations (90) and (95) we have 

N b 2 1000 Nb 2 v f> 

2.5 168,000 ’ 


from which it follows that 

v p = 67.2 pounds per square inch. 
which is the assumption on which the A. L. A. M. rating 
formula is based. The two formulae above discussed give 
widely different values when applied to motors of large bore 
and long stroke, but it must always be remembered that the 









POWER OUTPUT AND OTHER CHARACTERISTICS. 365 


first is intended to give the maximum power the motor can 
possibly deliver, whereas the second is designed to give the 
power at 1,000 feet piston speed per minute, which can now be 
exceeded by most motors, and greatly by motors of unusually 
long stroke. 

In 1912 the Technical Committee of the Automobile Club of 
France, in a report to the French Government, recommended 
horse power formulae based on the following brake mean effective 
pressures: Touring car motors running at not over 1,000 ft. p. m. 
piston speed, 75.2 lbs. p. sq. in.; high speed touring motors, 78.1 
lbs.; truck motors, 82.4 lbs. 

Ratio of Bore to Stroke—The first motors built for auto¬ 
mobiles for the most part had a comparatively long stroke, 
50 per cent, or more longer than the cylinder bore, automo¬ 
bile engine designers in this respect following the practice 
of stationary gas engine designers. Then came the era of the 
short stroke, high rotary speed engine, inaugurated by Bouton 
and Buchet in France. The short stroke motors of these en¬ 
gineers certainly weighed considerably less, horse power for 
horse power, than any previously produced, and so their ex¬ 
amples soon found many followers. Since that time, about 
1900, whenever it was desired to produce an engine of high 
weight efficiency, the stroke was made less or equal to the 
bore. For a time the opinion prevailed that the power output 
did not depend upon the bore at all; that a 4x4 inch motor, 
for instance, would develop as much power as a 4x6 inch 
motor, by giving its maximum output at a rotative speed 50 
per cent, higher, while, on the other hand, the power in¬ 
creased as the square of the bore or faster. However, this 
idea was exploded by the results of the so-called 4 inch or 
100 mm. races in Europe several years ago. In these races 
the cylinder bore was limited to 4 inches, and in every case 
the greatest speeds were developed by cars with long stroke 
motors, thus showing conclusively that the long stroke motors 
developed greater power than the short stroke ones. It may 
here be added that Charles Faroux, a French technical automo¬ 
bile writer, found that the output of a motor varies as the 0.6 
power of the length of stroke, and the German Automobile Tech¬ 
nical Association adopted a formula according to which it varies 
as the 0.7 power of the stroke. 

Effect of Length of Stroke on Thermal Efficiency—It hav¬ 
ing thus been shown that a long stroke motor will develop 


366 POWER OUTPUT AND OTHER CHARACTERISTICS. 


greater maximum power and at a higher piston speed than a 
short stroke motor, it remains to show the reasons therefor. 
One of these is that the long stroke increases the thermal 
efficiency. Consider two motors, both of 3 inch bore and of 
3 and 6 inches stroke, respectively. Let the compression ratio 
in each be 3. Then, considering the compression chamber 
as an extension of the cylinder, its wall area will be 


2 X 3 X 3 X 3.14 
4 


+ 3 X 3.14 X 1 = 23.55 square inches 


in the case of the short stroke motor, and 


2 X 3 X 3 X 3.14 
4 


+ 3 X 3.14 X 2 — 32.97 square inches. 


in the case of the long stroke motor. The wall area of the 
combustion chamber in the long stroke motor is, therefore, 
little more than 50 per cent, greater than that of the combus¬ 
tion chamber of the short stroke motor, while the charge 
drawn in and the heat developed by it is 100 per cent, greater. 
The advantage is on the side of the long stroke motor 
throughout the stroke, though it is greatest near the be¬ 
ginning of the stroke when the temperature difference on op¬ 
posite sides of the cylinder wall, and hence the heat loss, is 
the greatest. 

Effect on Volumetric Efficiency—Another reason is the 
smaller effect of inertia on the incoming charge in the long 
stroke motor. The speed at which charge flows into the cyl¬ 
inder under a given suction force depends partly upon the 
resistance to gaseous flow of the inlet passage and partly upon 
the inertia of the gas column filling this passage. Now, con¬ 
sider again the two motors mentioned above. If the piston 
speed were the same in each the suction stroke would be of 
twice the duration in the long stroke as in the short stroke 
motor, and the maximum and mean gas speeds through the 
inlet passage would be the same in both motors, on the supposi¬ 
tion that the volumetric efficiency is the same in both. But in the 
long stroke motor the gas would have twice the length of 
time for attaining its maximum speed, and it would, there¬ 
fore, be hampered much less by inertia. The result is that, 
since the inertia effect on the incoming charge is less, the long 
stroke cylinder will fill better than the short stroke cylinder, 
at any given piston speed. The same volumetric efficiency 
may therefore be obtained in the long stroke motor at a 
higher piston speed, and it is well known that decrease in 
volumetric efficiency is the chief factor limiting the working 
speed of gasoline motors. This is probably the most impor- 





POWER OUTPUT AND OTHER CHARACTERISTICS. 367 


taut reason why the long stroke motor runs at higher piston 
speed and develops greater power than the short stroke motor. 

A third reason is to be found in the reduced inertia effects 
on the reciprocating parts in the long stroke motor. We 
found (equation 24) that the inertia force is equal to 

/'a = 0.0000142 WIN 2 (cos 6 -(- cos 2 6 ) founds. 

2 n 

1 he crank angle factor in the above equation will be independent 
of the length of stroke if the ratio n of connecting rod length 
to length of stroke remains the same. Assuming the same 
weight of reciprocating parts in motors whose strokes are / 
and 2/ respectively, the inertia effects on the parts will be the 
same if 

/« 1 2 = 2 / w 2 2 , 

from which it follows that 

n i — 1 — n x = 0.707 

Since the stroke is twice as long and the angular speed 
0.707 times as great, the piston speed in the long stroke motor 
will be 1.414 times as great. However, the weight of the re¬ 
ciprocating parts will be somewhat greater in the case of the 
long stroke motor, but not as much as 41 per cent. 

Length of Stroke and Weight Efficiency—There can be 
no doubt that a long stroke motor is heavier than a short stroke 
motor of the same horse power output, all other things being 
equal. The three heaviest parts of any motor are the cylinders, 
the crank case and the flywheel. The weight of the cylinder 
varies almost as fast as the length of stroke, that of the crank 
case varies faster than the length of stroke, and the flywheel 
weight, for the same coefficient of irregularity, varies as the 
square of the length of stroke (equation 81). Hence the total 
weight of the motor varies probably as fast as the length of the 
stroke, while, according to. authorities cited, page 365, the 
output varies only as the 0.6 and 0.7 power of the stroke, re¬ 
spectively. It is very difficult to make an exact comparison, be¬ 
cause of the many parts involved, and no practical data on the 
subject are available, because cases of engines of different stroke- 
bore ratios designed along otherwise identical lines are very rare. 
Designers, however, show their belief in this principle by using 
short stroke motors wherever high weight efficiency is paramount, 
as in aeronautic motors. The great majority of modern aero¬ 
nautic motors have a stroke-bore ratio of one or slightly over, 
while in recent automobile engine practice abroad the ratio is 
much higher, especially in small bore motors. 

Service for Which Motor Is Intended Should Govern 
Ratio.—The present tendency is to give motors of small horse 




368 POWER OUTPUT AND OTHER CHARACTERISTICS. 


power output intended for town cars, taxicabs, etc., a stroke 
relatively long as compared with the bore, and more powerful 
motors for touring cars a stroke not so long in relation to the 
bore. There are two reasons for this practice. The first is 
that in a low powered car intended for slow speeds the problem 
of the gear reduction to the rear axle becomes easier if the 
motor turns at a relatively low angular speed—that is, if it is a 
long stroke motor—while in the case of a fast touring car no 
difficulty is experienced in getting the necessary reduction ratio. 
Take, for instance, the case of a town car with 32 inch wheels, 
to be driven at 25 miles per hour at the maximum output of the 
motor. At the above speed the road wheels will rotate at 265 
revolutions per minute. Now, suppose the motor to have a 
stroke of 4 inches. At a piston speed of 1,000 feet per minute 
it would turn at 1,500 revolutions per minute and the required 
reduction ration would be 

Ii5P_°- 5 .66. 

265 

It is difficult to obtain this ratio and still have a pinion with an 
adequate number of teeth of the proper pitch and sufficient clear¬ 
ance under the rear axle housing. It is therefore better to 
increase the length of stroke and decrease the speed of revolution 
•of the motor, thereby decreasing the required reduction ratio. 

If, on the other hand, we take a touring car with 36 inch wheels 
to be geared to 45 miles per hour at 1,000 feet piston speed of 
the motor, the speed of the rear axle will be 420 revolutions per 
minute. Then, for a reduction ratio of 3 we would get a 
length of stroke of about 5^ inches. Among the more recent 
motors for pleasure cars the stroke is seldom less than 5 inches, 
although the bore is sometimes as small as 3 inches. A diagram 
prepared by the Horse Power Formula Committee of the Insti¬ 
tution of Automobile Engineers, in which the stroke-bore ratios 
of a large number of motors are plotted against the bores, shows 
that the average ratio decreases from about 1.75 for 2.5 inch 
bore to nearly 1 for 5 inch bore. 

Dependence of Flexibility on Length of Stroke—There 

is a general impression that the long stroke motor is more 
flexible than a short stroke motor. By this is meant that the 
ratio of the “maximum load” speed to the minimum speed under 
full throttle is greater in the case of the long stroke than in 
that of the short stroke motor. 

We have already shown that the long stroke motor will give 
its maximum output at a higher piston speed, and if it can be 



POWER OUTPUT AND OTHER CHARACTERISTICS.' 369 

shown that it will run regularly at an equal or lower minimum 
piston speed under full throttle as the short stroke motor the 
point will have been proved. This is obviously largely a matter 
of flywheel capacity, but it depends also to some extent upon 
the valve setting. In a short stroke motor the inlet valves must 
have a large lag and the exhaust valves a great lead, in order 
that the motor may develop its maximum power at a relatively 
high piston speed. But this setting of the valves interferes with 
the operation of the motor at low piston speed under full throttle. 
In a long stroke motor less lag and lead are required in the valve 
action, and if such a motor is provided with a sufficiently heavy 
flywheel it may well be able to run at a lower piston speed under 
full throttle than a short stroke motor. However, both the speed 
of maximum output and the minimum speed under full throttle 
are dependent upon several interrelated factors, so that it is 
difficult to determine the direct effect of any one factor upon 
the flexibility. 

It may here be well to correct an erroneous impression often 
met with—that the hot gases are expanded farther in a long 
stroke than in a short stroke motor, and that in consequence 
the working efficiency is increased, the same as by increasing 
the ratio of expansion in a steam engine. This view is abso¬ 
lutely incorrect, since the ratio of expansion is the same as the 
compression ratio and is not at all affected by the length of 
stroke or by the stroke-bore ratio. 

Characteristic Curves—In addition to the maximum power 
that can be obtained from a motor, the horse power outputs at 
various angular speeds and the corresponding torques are of 
interest to the engineer. These values are preferably plotted on 
co-ordinate diagrams with the speeds of revolution as abscissas 
and the horse powers and the torques in pounds-feet as ordinates. 
A fuel consumption curve is frequently plotted on the same 
diagram. 

A set of such characteristic curves Is shown in Fig. 237. These 
were taken from a six cylinder Alco motor manufactured by 
the American Locomotive Co., Providence, R. I., of 4 Y\ inches 
bore by 5*4 inches stroke. The motor is of T-head type, with a 
compression volume to total volume ratio of 21.2 per cent. It 
has high tension magneto ignition, with spark plugs set over 
the inlet valves. It will be seen that the minimum speed at 
which observations were taken was 450 revolutions per minute. 
The torque increased slightly to 750 revolutions per minute, and 
then decreased, at first slowly, and then more rapidly. The 
maximum output of 64.5 horse power is obtained at 1,360 revo- 


370 POWER OUTPUT AND OTHER CHARACTERISTICS. 


lutions per minute. This corresponds to a mean effective pres¬ 
sure "n p of 64.2 pounds per square inch and a piston speed of 
1,247 feet per minute. The specific fuel consumption is also 
shown on the diagram. It will be seen that this is a minimum 
at about 900 revolutions per minute, viz., 0.76 pound per brake 
horse power-hour. The torque is a maximum at 750 revolutions 
per minute, and the mean effective pressure is then 81.6 pounds 
per square inch. 


65 

60 

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Fig. 237.— Characteristic Curves of Alco Six Cylinder, 

Inch Motor. 


Another set of characteristic curves are shown in Fig. 238. 
These are of a White four cylinder 1913 model motor, and the 
tests were made at the laboratory of Joseph Tracy, Rutherford, 
N. J. The motor "had a bore of 2 >Ya inches and a stroke of 5% 
inches. All valves are located on one side, and the motor is of 
the en bloc type. Ignition is by magneto. It will be seen that 







































POWER OUTPUT AND OTHER CHARACTERISTICS. 371 

this motor delivers its maximum output of slightly over 33 horse 
power at 1,600 revolutions per minute, corresponding to a piston 
speed of 1,366 feet per minute. The brake mean effective pres¬ 
sure at the point of maximum torque (about 1,000 revolutions 
per minute) figures out to 86 pounds per square inch, and at the 
point of maximum output (1,600 revolutions per minute) to 71 
pounds per square inch. The fuel consumption is rather low, 
only slightly over 0.6 pounds per horse power hour. Undoubt¬ 
edly, the carburetor was adjusted for high economy, as the test 
was made in the interest of a carburetor representative. 



Fig. 238.— Characteristics of Four Cylinder, Inch 

1913 White Motor. 


Power and torque curves of a Moline-Knight four cylinder 
sleeve valve motor (Chapter XVIII) are shown in Fig. 239. 
This motor, with a bore of 4 inches and a stroke of 6 inches, was 
subjected to a 336 hour non-stop test at the laboratory of the 
Automobile Club of America in December, 1913. In addition to 
the non-stop test, a number of output and fuel consumption tests 
were run, and the results of one of these tests are shown 
graphically in the figure. It appears that the maximum output 




































372 POWER OUTPUT AND OTHER CHARACTERISTICS. 


was attained at about 1,800 revolution, though the test was hardly 
carried far enough to show this definitely. The brake mean 
effective pressure at the speed of maximum torque (1,000 revo¬ 
lutions per minute) figures out to 95 pounds per square inch, and 
at the speed of maximum output to 77.5 pounds per square inch. 
It will thus be seen that in the Knight motor the brake mean 
effective pressure, both at the speed of maximum torque and that 



Fig. 239.—Characteristic Curves of Moline-Knight Four Cyl¬ 
inder, 4x6 Inch Sleeve Valve Motor. 

of maximum output, is approximately 10 per cent, higher than 
in the particular poppet valve motor whose performance is 
represented by Fig. 238. 

Maximum Power from Given Cylinder Volume and 
Motor Weight—There are a number of expedients in design 
which may be resorted to when it is desired to obtain the 
maximum output from a motor of given cylinder dimensions or 
of given weight. The motor must, of course, be so designed 
that it will run at a very high piston speed. In the first place, 
the valves must open directly into the cylinder, instead of into 
pockets at the sides of the cylinder head. This not only makes 





































POWER OUTPUT AND OTHER CHARACTERISTICS. 373 


the inlet and outlet passages shorter and somewhat more direct, 
but enables the combustion chamber to be made of a more nearly 
ideal form, with less surface area. The compression ratio, of 
course, must be as high as possible without danger from pre¬ 
ignition when the motor runs cotninuously under full load. The 
limiting ratio seems to be a little over 5, but this can only be 
used for relatively small bores. This brings up the matter of 
cooling. Energetic circulation of the cooling water and a 
relatively large radiator capacity are required. Tests show that 
within the usual temperature range the output is the greater the 
cooler the cylinder walls are. This is undoubtedly due to the 
fact that when the cylinder walls are relatively cool, the incoming 
charge is expanded less, and more charge is therefore taken in. 
On the other hand, the fuel consumption per horse power seems 
to decrease with an increase in the cylinder wall temperature. 

The valve opening areas must be very large, and the valves 
must open and close as rapidly as possible. The exhaust valves 
must not be siamesed, but instead a separate pipe should be run 
from each exhaust outlet to an expansion chamber at some 
distance from the cylinders—if the motor must exhaust through 
a muffler. But any kind of muffler absorbs some power, and 
for this reason the motors on racing cars, as a rule, exhaust 
directly into the atmosphere through short lengths of pipes. 

In determining the valve timing the inertia of the gases must 
be taken into account. The exhaust valves must therefore open 
very early—as much as 60 degrees of crank travel ahead of the 
dead centre—and the inlet valve close very late—as much as 45 
degrees past the dead centre. The exhaust valve should close and 
the inlet valve open about 15 to 20 degrees past dead centre. 
A motor with such a timing will not run very well at low speeds, 
and will, therefore, not be so flexible as one with the more 
usual timing, but this is the price that has to be paid for high 
maximum output. 

The reciprocating parts must be made as light as possible, so 
as to keep down the inertia forces and the friction losses result¬ 
ing therefrom. In the early part of this work the average 
reciprocating weight was stated to be 048 pound per square 
inch of piston head area. This relates to standard commercial 
practice. The weight can, however, be made very much less. 
In Fig. 240 is illustrated a 4-8 inch hydraulically forged steel 
piston, as made by a French firm, which weighs only s J / 2 pounds. 
It may be inferred that a similar 4 inch piston would weigh 
about 2 x / a pounds. The piston pin and two rings together will 
weigh about pound, and the connecting rod can be made to 


374 POWER OUTPUT AND OTHER CHARACTERISTICS. 

weigh not over 2 pounds. Hence the total reciprocating weight 
would be only 3.75 pounds, or about 0.3 pound per square inch 
of piston head area. 

G. A. Burls, in a paper read before the Institution of Auto¬ 
mobile Engineers (Great Britain), found that the weight of the 
reciprocating parts where cast iron pistons are used may be 


expressed by the equation 

m = 0.08 b 3 (1 + 0.15 r ) + 1.5 pounds .(96) 

and where pressed steel pistons are used by 
m — o.osb 3 (1 -f 0.15 r) + 1.5 pounds .( 97 ) 


The Motor as a Brake—When the spark is cut off and one 
of the lower gears is engaged, the gasoline motor acts as an 
excellent vehicle brake. Its braking resistance is partly due to 



mechanical friction in the engine and partly to the loss of energy 
in compressing and re-expanding combustible charge. The loss 
in the latter operation is due to surface friction of the passages, 
leakage of the charge when compressed, and to dissipation of 
heat generated by the compression of the charge, through the 
cylinder walls. When the motor is thus driven by the car it 
will draw in a charge of combustible gas during the first or 
suction stroke, which will absorb a slight amount of power. 
This charge will be compressed during the second stroke, whereby 
a considerable amount of power will be absorbed. During the 
third stroke, however, the compressed charge will expand again, 
whereby nearly all the power absorbed by the compression will 
be given out again. During the fourth stroke the chafge will be 















































POWER OUTPUT AND OTHER CHARACTERISTICS. 375 


forced out of the cylinder, with the expenditure of a small 
amount of energy. 

The braking effect can be considerably increased by opening 
the exhaust valve at the beginning of the third stroke and hold¬ 
ing it open during the whole of that stroke. The charge com¬ 
pressed during the second stroke will then escape through the 
exhaust valve at the beginning of the third stroke and not 
expend its energy on the piston. 

A series of tests to determine the braking effect of a gasoline 
motor were made by Dr. W. Watson and reported by him in a 
paper read before the Automobile Club of Great Britain and 
Ireland on February 15, 1906. The tests were made on a two 
cylinder 3^4 by 4 inch engine at the comparatively low speed of 
590 revolutions per minute. The results of the tests were as 
follows : 

Ratio of 

Horse Braking to Power, 
Power. Per cent. 


Power delivered at clutch by engine when working. 3.7 

Braking effect with throttle closed. 1.32 33 

Praking effect with throttle closed and compression 

cock open. 1.69 46 

Braking effect with exhaust valve prevented from 

opening . 1.60 43 

Braking effect with exhaust valve opened during 

third stroke. 2.40 65 


The tests also showed that the mechanical friction was 0.75 
horse power, and, of course, the difference between the braking 
effect under different conditions and the friction represents power 
used in compressing and moving the air. 

The braking effect can be still further increased by suitable 
modifications of the valve gear, so that a charge will be com¬ 
pressed in the cylinder during every up-stroke of the piston and 
allowed to escape at the end of that stroke. One method of 
accomplishing this is made use of on the Saurer motor truck. 
In the engine of this truck, during the first stroke air is drawn 
in through the inlet valve, during the second both valves are 
closed and the air is compressed, during the third the exhaust 
valve opens and allows the air to escape and the cylinder to be 
drawn full again, and during the fourth the exhaust valve, which 
ordinarily is open during this stroke, remains closed and the 
charge is compressed. After this the cycle is repeated, the 
charge compressed during the fourth stroke escaping through 
the inlet valve immediately upon the beginning of the down 
stroke. This system, of course, necessitates shifting of the 
exhaust cams. 






576 POWER OUTPUT AND OTHER CHARACTERISTICS. 


A system has also been devised in which the braking effect 
can be varied by the operator. Both the inlet and exhaust cams 
are shifted longitudinally and provided with extra cam bosses 
which come into action when the cams are moved into the brak¬ 
ing position. Fig. 241 shows the form of the exhaust cam used 
with this system, and Fig. 242 the form of the inlet cam. It 
will be seen that the exhaust cam, in addition to the regular 
cam boss A, is provided with two extra cam bosses BB at right 
angles to the former. The three cam bosses are suitably in- 



Fig. 241.—Exhaust Cam for Motor Brake. 


dined at their adjacent ends so that the cam roller may travel 
from the regular boss on to the extra bosses, and vice versa. 
The inlet cam also has two extra bosses DD at right angles to 
the regular boss C, and all of its bosses are also inclined at the 
ends. The two extra bosses on the inlet cam are sloped so that 
they will hold the inlet valve open during smaller or greater 
fractions of the up-stroke, according to the longitudinal position 
of the cam relative to the cam roller. 



Fig. 242.— Inlet Cam for Motor Brake. 












































POWER OUTPUT AND OTHER CHARACTERISTICS. 377 

When the exhaust cam is shifted in the direction of the 
arrow, so as to bring the extra cam bosses underneath the cam 
roller, the exhaust valve will be kept open during the first and 
third strokes of the motor; that is, during each down-stroke. 
During these strokes air will be drawn in through the exhaust 
valve. This air will be compressed during the up-strokes, and 
most of it will be discharged through the exhaust valve at* the 
beginning of the following down-stroke. Simultaneously with 
the exhaust cam the inlet cam is shifted in the direction indicated 
by the arrow. If the cams are shifted to the limit of their range 
of motion, the cam bosses DD are out of line with the cam 
roller, and the inlet valves remain closed. But in intermediate 
positions of the cams the two extra cam bosses DD will hold 



Fig. 243.—Motor Brake Deceleration Curves. 

the inlet valve open during a smaller or greater fraction of the 
up-strokes, thus allowing some of the air to escape from the 
cylinder before it is compressed, so reducing the work of com¬ 
pression. Experience has shown that with motors arranged as 
above described the braking effect is as much as 80 per cent, of 
the maximum power output. 

The motor is thus transformed into a brake by simply shift¬ 
ing the camshafts or the cams thereon. This should be done by 
means of the same lever by which the speed of the motor is 
controlled. Some deceleration curves of a car equipped with 


•Speed in Miles per -Hour 




































378 POWER OUTPUT AND OTHER CHARACTERISTICS. 

a motor brake of the type above described are shown in Fig. 243. 
1 he chief advantage of the motor brake over the regular friction 
brake is that it can be used for any length of time without 
undue heating and without excessive wear. Besides, it can be 
applied gradually, with little effort and without first removing 
the hand from the speed control lever. Figs. 241-3 are from Der 
Motorwagen, of Berlin. 



As a Source of Gas Pressure—Gas pressure can be used 
on an automobile for many purposes, as for operating a sig¬ 
naling device or a motor starter, for inflating tires, for feeding 
fuel, etc. A supply of gas under any pressure up to about 100 
pounds per square inch can easily be obtained from one or 
more cylinders of a motor. As regards the use of the gases 
for the inflation of tires, it should be pointed out that the 
products of combustion are not as efficient for this purpose as 
is atmospheric air, for the reason that the gases filter through 
rubber much quicker than does air. 

In order to obtain a supply of gas under pressure from the 
motor for such purposes as operating a motor starter, a check 
valve, similar to that shown in Fig. 244 at A is used. This fitting 
is generally screwed into the hole for the priming cock, and 
the priming cock A in turn is screwed into the nipple B at the 
bottom of the check valve, as shown. The central portion C 
of the check valve is flanged so as to prevent the valve from 
attaining an excessive temperature. The check valve proper, D, 
is of the ball type, the motion of the ball being limited by a 
stop formed integral with the screw closing the hole through 
which the ball is introduced. From this check valve a tube is run 
to the tank in which the gas is to be stored. At each explosion 

































































POWER OUTPUT AND OTHER CHARACTERISTICS. 379 


in the cylinder the ball valve will lift and allow a small amount 
of gas to pass, but as soon as the pressure in the cylinder drops 
the valve will close and thus trap the gas which has passed it. 
As already stated, a pressure of more than ioo pounds to the 
square inch can thus be obtained in the tank. 

For feeding gasoline from a tank located under the vehicle 
frame at the rear to the carburetor under the motor bonnet 
in front or to an auxiliary tank on the dashboard, a pressure of 
several pounds per square inch is sufficient. This pressure can 
be obtained from the engine by means of a reducing valve, as 
shown in Fig. 244 at B. This valve is usually connected to 
the exhaust manifold just beyond its connection to the engine. 
The gases first pass through a strainer A which prevents 
particles of carbon from passing into the fuel tank and causing 
trouble in clogging the spray nozzle. Next, the gases pass 
through the check valve B and then out through the tube C 
which leads to the fuel tank. In order to insure that the gas 
pressure will never exceed the predetermined maximum value 
a second spring-loaded valve D is provided, whose spring is 
so adjusted that the valve will open when the predetermined 
maximum pressure is exceeded, allowing some of the gases to 
escape and the pressure to be relieved. 

Weight of Motors—There is considerable variation in the 
weight of different motors of the same cylinder dimensions. 
On the whole, the higher grades of motors generally weigh less 
than cheaply built ones. If the wall thicknesses are made 
small the best grade of workmanship is required, and, besides, 
in many parts the use of high grade materials permits of 
making the cross-sections smaller. It is a most remarkable 
thing that the weight per horse power of explosion engines 
varies from several hundred pounds in stationary gas engines 
to as low as two pounds in flying machine engines. Automobile 
engines are intermediate between these two extremes in respect 
to weight. A fair approximation to the weight of motors de¬ 
signed in accordance with the rules laid down in this work 
may be obtained by means of the following equations: 

Weight of a single cylinder motor = 2.8 b 2 1 pounds ; 
of a double cylinder = 3.6 b 2 1 pounds; 
of a four cylinder = 5 b 2 1 pounds ; 
of a six cylinder = 7 b 2 1 pounds, 
b being the bore and / the stroke in inches. 

These rules apply to motors with aluminum crank cases. 


380 POWER OUTPUT AND OTHER CHARACTERISTICS. 


In order to reduce the weight per horse power, the first essen¬ 
tial is that the reciprocating parts be made as light as possible, 
so as to permit of high piston speed and consequently increased 
output. Other expedients for reducing the weight of motors 
are the use of sheet metal jackets, block cylinders, or cylinders 
so cast as to permit of their being bolted together to form a 
block; hollow crankshafts, sheet metal crank bases, cast steel 
flywheels of large diameter and tubular or I-beam construction 
for all parts subjected to compression and bending stresses. 


CHAPTER XVIII. 


SLEEVE, PISTON AND ROTARY VALVE MOTORS. 

Up to about 1908 the only form o£ valve that had proven 
commercially successful on gasoline motors was the poppet valve. 
Dr. Otto, the inventor of the four cycle gas engine, in one of his 
earliest machines used a sliding valve, and all the different forms 
of valves used in steam engines, including flat slide, piston and 
rotary valves, were at one time or another applied to gasoline 
motors, but they never proved successful in the long run. In 1908 
Charles Y. Knight, of Chicago, induced the Daimler Motor Co., 
of Coventry, England, to take up a motor of his invention in 
which the valve functions are performed by two ported sleeves 
one within the other, and both between the piston and the cylinder 
wall. The two sleeves are reciprocated by a half speed shaft 
carrying small cranks equal in number to the sleeves. This 
motor has since proven very successful, and has been taken up by 
leading manufacturers in different European countries and in the 
United States. 

The Knight Motor—Fig. 245 is a sectional view through one 
of the cylinders of a Knight motor built by Panhard & Levassor, 
of France. Within the cylinder bore there is a thin sleeve A of 
cast iron, cut with ports on opposite sides for the inlet and ex¬ 
haust. Within this sleeve there is another, B, also cut with 
ports near its upper end, and within the latter sleeve is located 
the piston. In these motors the cylinder head C is detachable, 
and is of rather unusual construction. A part of it extends into 
the cylinder for quite a distance, but is of a diameter smaller 
than the cylinder bore, so the two sleeves may enter between it 
and the cylinder wall. Both the cylinder head and the piston 
head are concave or bowl shaped, the object being to secure as 
nearly as possible a spherical combustion chamber. The two 
sleeves are reciprocated by means of a half speed shaft, which 
is driven from the crankshaft through a silent chain. The half¬ 
speed shaft is formed with eight cranks, and these cranks con¬ 
nect through short connecting rods EE with lugs formed on the 


381 



382 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 


valve sleeves at their lower ends. The connecting rods for the 
two sets of sleeves are of different lengths, and usually they are 
slightly inclined, that is to say, the vertical line along which 
their upper end reciprocates does not pass through the axis ot 
the half speed shaft, but is somewhat closer to the cylinder axis— 
this for the sake of compactness. It will be noticed that the 
drive of the sleeves is unsymmetrical, a feature the practicability 
of which was questioned at one time, but it has been found that 



Fig. 245.—Section Through Cylinder of Panhard- 

Knight Motor. 

owing to the long bearing of the sleeves it does not cause any 
trouble. At the lower ends, where the driving lugs are, the 
sleeves are suitably strengthened by circumferential flanges. 

One peculiarity of the Knight motor is that the bore of the 
cylinder casting is not the effective bore of the engine, the latter 
being equal to the bore of the inner valve sleeve. The sleeves 












































































SLEEVE, PISTON AND ROTARY VALVE MOTORS. 383 


are made of a thickness equal to about 0.040 inch per inch of 
effective bore. The outer sleeve is sometimes made slightly thin¬ 
ner than the inner one, because it is not subjected to the pressure 
of the explosion. Thus, for instance, in the 100x140 mm. Pan- 
hard-Knight motor the inner sleeve is made 5 mm. and the outer 
sleeve 4 mm. thick. The bore of the cylinder casting in this motor 
is therefore 118 mm. 

Packing Ring Arrangement—Gas tightness is insured by 
two sets of packing rings, one set of three or four being placed 
on the piston in the usual way, and the other set on that portion 
of the cylinder head which extends into the cylinder and is sur¬ 
rounded by the valve sleeves. The latter set comprises one un¬ 
usually broad ring, F, at the bottom, referred to as a junk ring, 
and two or three rings of the same dimensions as used on the 
piston. The junk ring itself does not possess sufficient spring 
force to apply it properly against the valve sleeves, and spring 
rings are placed underneath it. At least, this is the construction 
used in most Knight motors. The object of the junk ring is to 
seal the ports in the valve sleeves while the explosion takes place 
and the gases expand in the cylinder. This ring, therefore, must 
be made of greater width than the height of the port. It is 
usually made about half an inch wider than the exhaust port. 
In Fig. 245 the moving parts are shown in the position corre¬ 
sponding to the beginning of the power stroke, and it will be 

seen that both the inlet and exhaust ports in the inner sleeve 
are fully covered by this junk ring. 

Port Dimensions—Each port extends substantially one-third 
around the cylinder. The ports are provided with a bridge at 
the middle of their length so as to prevent undue weakening of 
the sleeve. On the theory that the port area should be propor¬ 
tional to the piston head area, the height of the ports should be 

proportional to the effective bore or piston diameter. Most of 

the Knight motors built so far have been of about 4 inches bore 
and slightly over 5 inches stroke, and in these the inlet port is 
made inch and the exhaust port 54 inch high. 

Valve Motion and Timing—The valve action and the effect 
of various factors upon the valve timing can best be studied by 
means of the diagrams Figs. 246 and 247. It may be stated 
that the small cranks for driving the two valve sleeves are in 
practice set so as to make an angle of 60 to 90 degrees with 
each other. In the diagrams it is assumed that they are set at 
an angle of 80 degrees. It is further assumed that the half 
speed shaft is directly undern-earh the hinge joint to the lug of 
the sleeve, and that the connecting rods are of twice the length 


384 SLEEVE, PISTON AND ROTARY VALVE MOTORS 

of the valve motion. We have here a regular crank and con¬ 
necting rod mechanism, the same in every particular as the main 
crank and connecting rod of the engine, and the laws of motion 
previously deduced for this mechanism therefore apply in this 
case. We found (equation 20) that the distance which the 
piston moves from its topmost position, while the crank turns 
through an angle 9, is 

__ l , a\ \ j (sin 2 6\ 

x = — (1 — cos 0) + n 11 - I > 

2 \ 8 n 2 ) 

where n is the ratio of connecting rod length to length of stroke. 
In case the lug joint is offset from the vertical line through 
the half speed shaft by a distance f, we may apply equation (62), 



according to which the vertical distance between the axis of the 
half speed shaft and the centre of the lug on the valve sleeve is 

_ ; nil sin 0 f \ o 1 l 

x x ~ n l — — (- — I— ] 2 —cos 0 . 

2 \ 2 n n l] 2 

The curves in Figs. 246 and 247 were drawn by means of the 
first of these equations. The sleeve travel was assumed to be 
1 l /s inches. It will be noticed that in each diagram there are 
four curves, representing the up and down motion of the top 
and bottom edges of the inlet and exhaust ports in the sleeves 
respectively. In Fig. 246 there are also two horizontal lines 
representing the top and bottom edges, respectively, of the inlet 
port in the cylinder wall. The inlet begins to open when the 
bottom edge of the port in the outer sleeve, moving downwardly, 
passes the top edge of the port in the inner sleeve moving down- 


























SLEEVE, PISTON AND ROTARY VALVE MOTORS. 385 


wardly. The inlet closes when the bottom edge of the port in 
the inner sleeve, moving upwardly, passes the top edge of the 
port in the outer sleeve also moving upwardly. By reference to 
the scale at the bottom of the diagram it will be seen that the 
inlet extends over 200 degrees of crank motion. Now, it is 
customary to open the inlet 10 to 15 degrees after the crank¬ 
shaft passes the upper dead centre and to close it about 30 to 
35 degrees after the crankshaft passes the lower dead centre. 
Since the total opening extends over 200 degrees, or 20 degrees 
more than a half revolution, it is evident that the lag of inlet 
closing must exceed the lag of inlet opening by 20 degrees. We 



Crankshaft Motion 
Fig. 247. —Diagram of Exhaust Action. 


will therefore let the inlet open 10 degrees late and close 30 de¬ 
grees late. The point on the horizontal axle at which the inlet 
begins to open therefore corresponds to a crankshaft motion of 
10 degrees from the top dead centre. This enables us to plot 
the four dead centre positions on the horizontal scale, and to 
properly divide the whole scale. 

Fig. 247 is a diagram of the exhaust port action. The ex¬ 
haust begins to open when the lower edge of the port in the 
inner sleeve, moving downwardly, passes the lower edge of that 
portion of the cylinder head which extends into the cylinder. 


'Exhaust Port. • 





















386 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 

The exhaust closes when the top edge of the port in the outer 
sleeve, moving downwardly, passes the lower edge of the port 
in the cylinder wall. In order to secure the proper length of 
valve opening the ports in the inner and outer sleeves are 
placed at different levels, the one in the outer sleeve being at a 
lower level (^4 inch lower in the diagram). The exhaust 
opening under the assumptions here made extends over a period 
corresponding to 240 degrees of crank motion. The exhaust 



Fig. 248. —Inlet Timing of Stearns-Knight Motor. 


begins to open 50 degrees ahead of the lower dead centre and 
closes 10 degrees past the top of dead centre. In order to effect 

this valve timing the half speed shaft must be so set that when 

the motor crankshaft is in the top dead centre position at the 
beginning of the inlet stroke, the crank operating the inner 

sleeve has to travel through 15 degrees until it reaches the bot¬ 
tom dead centre position, and the crank operating the outer 

sleeve has to travel through 95 degrees till it reaches the bot¬ 
tom dead centre position. 



















































































SLEEVE, PISTON AND ROTARY VALVE MOTORS. 387 


Diagrams like the above are very convenient for studying 
the effect of various changes in the ports and valve gearing, as, 
for instance, a change in the angle between the cranks for the 
two sleeves of each cylinder, a change in the width of ports, 
etc. It may be pointed out that the various licensees under the 
Knight patent design their engines in their own way and employ 
different valve timings. As a rule, since the large area of port 
opening allows of very high piston speeds, the inlet valve is given 
a considerable lag and the exhaust port considerable lead. 



Fig. 249.— Exhaust Timing of Stearns-Knight Motor. 


Figs. 248 and 249 show the positions of the valves and the 
crank of a Stearns-Knight motor when the inlet begins to open, 
is fully open and closes, and when the exhaust begins to open, 
is fully open, and closes. This motor, contrary to ordinary 
practice, turns in a counter clockwise direction. It will be no¬ 
ticed that the exhaust has the extreme lead of 64^ degrees, 
but no lag. 






















































































388 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 

Port Opening Areas—The two chief advantages claimed for 
the Knight motor are, first, its high output for given cylinder 
dimensions, which is due partly to the large area and rapid 
opening and closing of the valve ports, and partly to the fa¬ 
vorable form of the compression chamber; and, second, its si¬ 
lent operation. It may be of interest to compare the valve 
capacity of a Knight motor with that of a poppet valve motor. 
Suppose a motor of 4 inch bore to have inlet ports inch 
high and extending one-third around the circumference. Also, 
let the valve timing be that depicted in Fig. 246. Then the 
opening of the valve is represented by the diagram Fig. 250. 
This diagram is practically a triangle, the opening curve being 
very slightly concave and the closing curve convex (from the 
inside). The average height of opening is therefore % inch. 
The length of the port around the circumference is 



4 . X 3 • *4 — 4 , inches , 

3 

and the valve capacity therefore is 

4.19 x V\ x 200 = 209.5 square inch-degrees. 

Now let us see what diameter of poppet valve would be required 
to equal this. Assume that the lift of the poppet valve is one- 
fifth of the clear diameter. Also assume that the valve is opened 
with constant acceleration and deceleration in 85 degrees of crank¬ 
shaft motion, and that the total opening period is 200 degrees of 
crank motion. Then the equivalent duration of full opening is 

200—85 = 115 degrees, 

and the full opening area of the valve must be 

2 _ ? . 9‘ _ 5 — x , 82 square inches. 

115 







SLEEVE, PISTON AND ROTARY VALVE MOTORS. 389 


Now, according to equation (88) the area of opening of a pop¬ 
pet valve is 

A = ir (o. 7 d h -f- o. 35 h 2 ) 

and with 

, __ d 
h — —> 

5 

A = 7r (o. 14 d 2 + 0.014 ^ 2 ) — 0.484 d 2 . 

We have, therefore, 

0.484 d 2 = 1 .82 square inches . 
d ~\/ 1 ' -8 2 =1.94 inches. 

y 0.484 

A diagram showing the variation of the opening area of a pop¬ 
pet valve of 1.94 inches clear diameter is superposed on the 
sleeve valve diagram in Fig. 250. The two diagrams should be 
of the same area. It appears that the sleeve valve diagram has 
a slightly greater area, which is probably due to a slightly er¬ 
roneous assumption in making the calculation. It is absolutely 
correct that the average lift of the poppet valve for the whole 
period of opening, viz., 200 degrees, is equivalent to the maxi¬ 
mum lift during 115 degrees of opening, but since the area of 
opening is not absolutely proportional to the lift (see equation 
88) the average opening of the valve for the full period is 
equivalent to a full opening for a slightly different period. The 
difference is negligible, however. 

It is frequently said that the sleeve valve opens much quicker 
than the poppet valve, but this is only partly correct. The fact is 
that the sleeve valve opens and closes at substantially uniform 
speed, whereas the poppet valve begins and ceases to open at 
practically no speed and attains its maximum opening speed at 
half opening. The sleeve valve opens the most rapidly at first 
but the least rapidly later on, and the poppet valve attains its 
full opening in a shorter time than the sleeve valve. 

Fig. 251 shows the opening diagram of the exhaust ports of a 
sleeve valve, the maximum width of the port opening being y 2 
inch, the length 4.19 inches and the timing as in Fig. 247. It will 
be seen that the port attains its maximum opening in 103 degrees 
of crankshaft motion, remains fully open for 27 degrees and closes 
240 degrees of crankshaft motion after beginning to open. The 
valve capacity, therefore, is 
2 

-X X 4.19 = 280 square inch-degrees. 

In order that a poppet valve having a lift equal to one-fifth the 
clear diameter, which is accelerated and decelerated uniformly 





390 SLEEVE, PISTON AND ROTARY VALVE MOTORS 

and attains its full lift in 85 degrees of crank motion, may have 
an equal valve capacity, it must have a diameter d such that 

0.484 d 2 = —— = 1.807 square inches 
240 — 85 

d — a / 1 - = 1.93 inches, 

y 0.484 

It may be said that comparatively few poppet valve motors are 
fitted with such large valves. 

Silent Operation— The other advantage of the sleeve valve 
motor is its silent operation at all speeds. This is due to the 
fact that the valves are closed (as well as opened) positively. 
Besides, the half speed shaft is driven by a silent chain instead of 
spur or helical gears. The absence of valve pockets and exposed 
valve gear members also gives the motor a simple and rugged 
appearance. 



Fig. 251.— Area-of-Exhaust-Opening Diagram. 

Power Consumed by Sleeves —Owing to the fact that most 
of the time while there is considerable pressure in the cylinder 
the sleeves travel in the same direction as the piston, compara¬ 
tively little power is required for moving the sleeves. It can be 
seen by reference to Fig. 246 that both sleeves move up with the 
piston during the compression stroke and that the inner sleeve 
moves down with the piston during the power stroke. The outer 
sleeve, however, moves in opposition to the piston during the 
power stroke. Mr. Knight states that tests made on a six cylin¬ 
der, 75 h. p. motor showed that 2 h. p. was required for driving 
the half speed shaft. 

Lubrication and Cooling —One of the chief problems with 
a motor of this kind is undoubtedly effective lubrication of the 












SLEEVE, PISTON AND ROTARY VALVE MOTORS. 391 


sleeves. Most of the motors manufactured under the Knight 
patents are provided with splash lubrication, and it is claimed that 
the suction in the ports during the inlet stroke draws the oil up 
between the sleeves, the piston and the cylinder wall. The 
splashers secured to the connecting rod heads dip into oil troughs 
underneath the cranks, which are hinged and connected with the 
throttle so that they are raised and lowered as the throttle is 
opened and closed respectively. This is illustrated in Fig. 252. In 
the Panhard-Knight motor, however, oil is fed to the sleeves 
through a passage in the cylinder head. The sleeves in every 
case are cut with helical oil grooves on the outer surface, and 
in the Panhard motor thev are in addition drilled with numerous 



Fig. 252.—Movable Oil Splash Trough. 


* * * 

oil holes so that the oil may pass from one side of the sleeve 
wall to the other. 

The question has been raised as to whether there would 
not be danger of the edges of the ports being burned by the 
hot gases, but it is stated that inasmuch as both of the ports 
in the inner sleeve are covered by the junk ring at the moment 
of explosion and for some time thereafter, their edges are not 
exposed to the maximum temperature occurring in the cylin¬ 
der, and no trouble from burning is experienced. 

Test Results of Knight Motor—A very rigid test of a 
four cylinder Knight motor of 96 mm. bore and 130 mm. 
stroke was made during March, 1909, by the Royal Automo¬ 
bile Club of Great Britain and Ireland. At first a bench test 
was made extending over five days twelve hours and fifty- 
eight minutes. No penalized stops occurred, but two stops of 
an aggregate duration of seventeen minutes were made which 
incurred no penalty. According to the conditions of the test, 
35.3 h. p. was to be obtained at 1,400 r. p. m., and the power 





















392 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 

was at no time to fall below this figure. The average horse 
power recorded during the time of the trial was 38.83 (A. L. 
A. M. rating 22.85). The gasoline consumed amounted to 
476.5 gallons, equivalent to 0.668 pound per horse power hour. 
On the completion of this bench test the engine was fitted 
into a chassis, under observation, without any of the vital 
parts being disturbed, and the car was fitted with a standard 
touring body. The car was then driven 229 miles on the road 
and 1,915.1 miles on Brooklands track, the average speed made 
on the track being 41.88 miles per hour. The gasoline con¬ 
sumption on the track was at the rate of one imperial gallon 
per 22.44 miles, equivalent to 33.37 ton-miles per imperial 
gallon. At the conclusion of the road test the engine was 
removed from the chassis, and, without having any of its vital 
parts disturbed, was submitted to another bench test which 
lasted for five hours two minutes. During this test there 
were no stops whatever, the average horse power recorded 
was 38.96, and the average gasoline consumption 0.677 pound 
per horse power hour. The engine was then completely disman¬ 
tled and no perceptible wear was noticeable.—A 336 hour non¬ 
stop run under full throttle at 1,100 feet piston speed per minute 
was made by a Moline-Knight engine at the Automobile Club of 
America laboratory in December, 1913 (see page 372). 

Piston Valves—Since the Knight motor proved to be a 
success there has been considerable development in gasoline 
engine valves of other than the poppet valve. A large number 
of patents have been taken out, particularly in England, for 
piston and rotary valve motors. The piston valve, as a rule, 
consists of a piston similar to the one in the working cylin¬ 
der which reciprocates in a valve cylinder generally parallel 
with the cylinder, but which may extend horizontally across 
the cylinder head or open into the cylinder head at an angle. 
This valve piston is driven from a half speed shaft through 
a small crank and connecting rod. In some cases the valve 
cylinder opens directly into the working cylinder, while in 
other cases it communicates with it through a port. In the 
former case there is a passage in the side of the valve cylinder 
which is uncovered by the valve piston during the desired 
period of valve opening. A design of the second kind is 
shown in Fig. 253. Here the valve piston is double ended, 
with a partition wall in the centre and a port in its side wall 
just above the partition wall, which registers with the port in 
the wall of the working cylinder at the proper time. Since 
the port in the valve piston registers with the cylinder port 


SLEEVE, PISTON AND ROTARY VALVE MOTORS. 393 

during both the up and the down stroke of the valve pistons, 
the valve operating shaft in this instance must be geared to 
run at one-quarter crankshaft speed. It is obvious that with 
a valve construction of this kind the valve diagram is of sub¬ 
stantially triangular form, the cylinder port being uncovered 
and then covered at practically uniform speed. The objection 
to the design is that it is rather bulky. 


JITxhaU'St /nlet 



Rotary Valves—Numerous designs of rotary valves have 
also been worked out, but they have not yet been adopted for 
automobile work to any extent. Fig. 254 shows the Reynolds 
rotary valve motor made by the Reynolds Motor Co., of De¬ 
troit, more particularly for marine service. The valve of this 
motor consists of a flat disc which seats against the inside of 
the cylinder head. It is provided with a shaft extending up 
ward through the centre of the cylinder head, and provided 



















































394 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 


with a spur gear on top of the cylinder head through which 
it is driven from the crankshaft at one-half the speed of the 
latter. The valve disc has a port cut into it of the shape 
shown in Fig. 255, which registers successively with the inlet 
and exhaust ports of the same shape in the cylinder head. 
The valve mechanism of this motor is quite noiseless and the 
motor is of simple and compact design, in which respect it 
resembles the sleeve valve motor, but it has the disadvantage 
that the valve is pressed against the seat by the force of the 
explosion, which causes considerable resistance to its motion, 
produces heat and tends to squeeze out the lubricating oil be¬ 
tween the valve disc and its seat. One advantage of this 
design of valve is the form of the ports, which are nearly 



Fig. 254.—Reynolds Rotary Valve Motor. 

































































































































































































































SLEEVE, PISTON AND ROTARY VALVE MOTORS. 395 



square and have a small periphery as compared with their 
area. This means that the resistance of the port is relatively 
small. 

Another form of rotary valve motor, the Mead, is illustrated 
in Fig. 256. In this four cylinder motor there are two long 
valve drums extending along opposite sides of the cylinders 



Fig. 256.—Mead Rotary Valve Motor. 












































396 SLEEVE, PISTON AND ROTARY VALVE MOTORS. 


near their heads. These valve cylinders are cut with diame¬ 
tral ports through them at distances apart equal to the cen¬ 
tre distances of the cylinders, and suitably spaced angularly 
so as to open the ports of the cylinders in the succession in 
which these cylinders fire. One of the valves serves for the 
exhaust and the other for the inlet. These valves are driven 
at quarter engine speed. 

The subject of piston and rotary valves for automobile mo¬ 
tors is as yet in its infancy. Those interested in it will do 
well to look up articles by Eugene P. Batzell in The Horse¬ 
less Age of May 25, June 1 and June 8, 1910, and September 
20, September 27 and October 4, 1911. 


CHAPTER XIX. 


AIR COOLING. 


The characteristic curves Figs. 237 and 238 show that a 
gasoline motor under the most favorable conditions con¬ 
sumes about 0.8 pound of gasoline per brake horse power- 
hour. In Chapter I it was stated that gasoline contains about 
19,000 B. T. U.’s per pound. Since one horse power is equal 
to 33,000 foot-pounds per minute, one horse power-hour is 
equal to 

60 x 33,000 = 1,980,000 foot-pounds. 

Also, since one B. T. U. is equal to 772 foot-pounds, the heat 
energy of 0.8 pound of gasoline is 

0.8 x 19,000 x 772 = 11,734,400 foot-pounds. 

Hence 


I, 980,000 

II, 734,400 


= 17 per cent. ( appr .) 


of the fuel energy is converted into useful work. Of ‘course, 
some energy is lost in friction at the motor bearings, in driv¬ 
ing the camshaft and accessories, etc., and we will not be 
far wrong in assuming that this loss amounts to about one- 
third of the brake horse power, so the proportion of the heat 
energy of the gasoline converted into mechanical energy is 
approximately 22 per cent. This leaves 78 per cent, which is 
either absorbed by the cylinder walls or discharged with the 
exhaust. 

As to the amount of heat discharged with the exhaust, the 
following computation may give a fair idea. We found that 
for best economy the gasoline must be mixed with about 
eighteen times its weight of air. The products of combus¬ 
tion of 0.8 pound of gasoline, therefore, weigh 

(18+1)0.8 = 15.2 pounds. 

Experiments show that the temperature of the exhaust gases 
on leaving the cylinder is about i,ioo° Fahr. Assuming the 
initial temperature of the charge to be 6o° Fahr., the excess 
temperature of the exhaust gases is 1,040° Fahr. We may 


397 




398 


AIR COOLING. 


take the specific heat of the exhaust gases to be 0.20. The 
heat energy contained in the exhaust gases is therefore 
15.2 x 1,040 x 772 x 0.20 = 2,440,000 foot-pounds. 

This is equivalent to 

■ 2 » 44 °»°gg_ = 21 per cent. 

ir,734,400 

If 22 per cent, of the fuel energy is converted into mechanical 
work and 21 per cent, is discharged with the exhaust, then 57 
per cent, must be absorbed by the cylinder walls. Tests made 
on water cooled motors usually show that about one-half of 
the fuel energy is taken up by the cooling water. 

Why Cylinders Must Be Cooled—Since such a large pro¬ 
portion of the heat of combustion is absorbed by the cylin¬ 
der walls, they must be subjected to very energetic cooling 
action if they are to be maintained at a temperature where lu¬ 
brication is still possible. Undoubtedly in the earliest gas en¬ 
gines no attempt was made to surround the cylinder with a 
water jacket for carrying off the heat, but it was probably 
very soon found that a plain cast iron cylinder when sub¬ 
jected to the heat of burning hydrocarbon gases on the inside 
quickly reaches such a temperature that mineral oils of even 
the highest flash point will burn on the walls and that 
lubi^ation will then cease. Besides, as the cylinder walls 
become very hot, the fresh charge upon entering the cylinder 
is unduly heated and expanded, and hence a smaller weight of 
charge enters the cylinder. Another trouble that is encoun¬ 
tered if the cylinder is inadequately cooled is that during the 
compression stroke the charge may reach such a high tem¬ 
perature as to ignite spontaneously, thus producing a violent 
shock on the piston and connecting rod while these parts are 
traveling in a direction opposed to the gaseous pressure on 
the piston head. Moreover, if the cylinder walls are allowed 
to attain a very high temperature they, as well as the sur¬ 
rounding parts, will be distorted, which causes increased fric¬ 
tion, leakage, etc. 

Air cooling was used by Gottlieb Daimler in his first bi¬ 
cycle motor, but was soon discarded by him. It was suc¬ 
cessfully applied to cycle motors by Bouton about 1897. 
These motors had single cylinders of very small bore and 
stroke. It is easily proven that a small cylinder can be kept 
cool better than a large one. With the same temperature of 
cylinder wall the amount of heat radiated per unit of wall 
area will be the same irt cylinders of all sizes. But since in a 
large cylinder the wall area is smaller, in proportion to the 



AIR COOLING. 


399 


combustion chamber volume, more heat will have to be ab¬ 
sorbed per unit of wall surface in order that the burning gases 
may drop in temperature as rapidly as in the small cylinder. 
Moreover, in some air cooled cycle or vehicle motors one side 
of the cylinder is exposed to a current of air which is either 
due to the motion of the vehicle or is produced by a fan. The 
other side of the cylinder, however, is shielded from this 
draught, and this shielding is the more effective the larger 
the cylinder dimensions. Consequently the tendency to over¬ 
heat is the greater the larger the cylinders. 

Cooling Flanges.—The amount of heat which will be ab¬ 
sorbed or radiated by a unit of metallic surface depends upon 
the temperature difference between the surface and the sur¬ 
rounding air or gas. If the combustion chamber wall had the 
same area of surface in contact with the hot gases on the in¬ 
side and the atmospheric air on the outside, then it would 
tend to assume a temperature which would be a mean be¬ 
tween the atsmospheric temperature and the average tempera¬ 
ture inside the cylinder, particularly if the conditions of motion 
of the gaseous elements inside and outside the cylinder, with 
respect to the cylinder walls, were about the same. But this 
temperature would be far too high to permit of satisfactory 
lubrication, not to mention other likely troubles. According to 
some experiments made by Prof. H. L. Callender, the operating 
conditions of an air cooled motor cease to be satisfactory when 
the cylinder head attains a temperature of 570° Fahr. 

Reduction of Heat Absorption—In order to keep down 
the cylinder wall temperature it is necessary to so shape and 
finish the inner surfaces that they will absorb a minimum 
amount of heat, and the outer surfaces so that they will radi¬ 
ate this heat most readily. It has been found by laboratory 
experiments that a polished metal surface does not radiate 
heat nearly as rapidly as a rough surface, and it is logical to 
assume that when in contact with hot gases it will not absorb 
as much heat either. It is therefore advisable to give a 
smooth machine finish to the whole interior surface of the 
combustion chamber. This can only be done if the valves 
are located in the cylinder head, and this valve arrangement 
is therefore a favorite one for air cooled motors. There is 
one other advantage in polishing the combustion chamber 
wall surface. It is found that spontaneous ignition or pre¬ 
ignition in both water and air cooled motors is often caused 
by carbon deposits on the cylinder walls. Carbon is a poor 


400 


AIR COOLING. 


conductor of heat and attains a very high temperature when 
adhering to the combustion chamber wall. When the wall 
surface is polished the carbon formed by the combustion will 
not so readily adhere to it, but will be blown out with the 
exhaust. 

The temperatures of explosion and combustion depend di¬ 
rectly upon the compression pressure. In an air cooled en¬ 
gine the compression, therefore, must be kept fairly low. A 
compression ratio of 4 to i is about the practical limit for air 
cooled motors, while a ratio of 5 to 1 is occasionally used 
on water cooled motors. Rich mixtures tend to heat the motor 
more than mixtures properly proportioned or somewhat lean, 
and it is, therefore, advisable to so arrange the carburetor that 
it is impossible for the driver to misadjust it so as to make 
the mixture over-rich. Quick evacuation of the cylinder at the 
end of the power stroke is also essential, and large size, 
quick opening exhaust valves must therefore be used. The 
H. H. Franklin Manufacturing Company has employed auxiliary 
exhaust ports, as already described, for many years, but has 
recently given up their use. These auxiliary ports open much 
faster than ordinary poppet valves. It is also inadvisable to 
run on a late spark. 

Radiating Surface—The combustion chamber wall disperses 
the heat which it absorbs from the burning gases by two 
methods, viz., by radiation and by convection. Suppose a cur¬ 
rent of air to be blown over the cylinder. Then any particle 
of air, on coming in contact with the hot cylinder wall, will 
have heat imparted to it and will carry this heat away with it. 
This is known as convection. There would also be convec¬ 
tion of heat if no artificial air current was employed, since 
the air in contact with the hot walls, on becoming heated, ex¬ 
pands and rises, thus carrying heat away with it. But there 
is some loss of heat from the cylinder walls at right angles 
to the direction of the air current or even directly opposed 
to it, and this is due to radiation. Convection is proportional 
to the speed at which the air moves over the hot surfaces. 
Both radiation and convection depend upon the area of the 
surface exposed and to some extent upon the condition of 
this surface. It is therefore necessary, in order to effectively 
cool the combustion chamber, to make the outer surface of 
the cylinder wall as large as possible and to pass cool air 
over it at a rapid rate. 

Form and Dimensions of Cooling Flanges—In very small 
motors, such as cycle motors, the only provision made for in- 


AIR COOLING. 


401 


suring effective cooling consists in increasing the radiating 
surface by providing the cylinder with cooling flanges or ribs. 
In a vertical motor these generally lie in a plane perpendicu¬ 
lar to the cylinder axis, as shown in Fig. 257. In the smaller 
motors these ribs are generally cast integral with the cylin¬ 
ders. A mathematical investigation of the form insuring the 
most effective radiation shows that the flanges should be of 
what might be called razor blade section; that is, slightly con¬ 
cave on their sides; should run out into a knife edge, and 
there should be a sharp angle between adjacent flanges where 



Fig. 257. —Cylinder With Cooling Flanges. 

they join the cylinder wall. Practical considerations dictate 
a somewhat different form, however (see Fig. 258). The 
section of the flange tapers down from the cylinder wall to 
the outer edge, about % inch per inch. This is required in 
order to make it possible to draw the cylinder pattern out 
of the mold, but it also possesses advantages from the stand¬ 
point of cooling, aside from approximating the shape of 
maximum radiating efficiency as just explained. Not as much 
heat needs to be conducted through sections of the flange 
near the outer edge as farther inward, and the section there- 



























402 


AIR COOLING. 


fore can be reduced as the distance from the cylinder wall 
increases. The bottom of the groove between adjacent flanges 
is generally made semi-circular and the outer portion of the 
flange the same, in accordance with ordinary rules of design. 

There is a considerable variation in design, but in small 
cylinders, say, not exceeding a bore of 3 inches, to which this 
form of cooling is now practically limited, the flanges are 
made about inch high per inch of bore and spaced to 
inch. Frequently the flanges are made highest at the top end 
of the cylinder and of gradually decreasing height toward the 
bottom end. 

Instead of casting cylinders with cast iron cooling flanges, 
sheet metal flanges are sometimes placed in the mold and cast 
into the cylinder walls. Again, the cylinders may be cast 
with cooling pins, sheet metal flanges may be slipped over 
the machined cylinder or threaded radiating pins may be 

screwed into its walls in large number. 
All of these methods were applied at one 
period in the history of automobile de¬ 
velopment, but, owing to a seeming re¬ 
version of public favor to the water cooled 
motor, many manuufacturers who once 
built air cooled motors gave up air cool¬ 
ing. As the power output demanded by 
the buying public increased the problem 
of cooling the motors satisfactorily by air directly became 
more difficult, and, besides, the fact that four cylinder motors 
arranged vertically in a fore-and-aft direction under a bonnet 
in front became the favorite type introduced further com¬ 
plications. It was at first attempted to cool these motors by 
a blast of air created by a propeller type of fan located at the 
forward end of the bonnet, the blast being directed against 
the front cylinder. Under these conditions the different cyl¬ 
inders naturally were cooled very unequally, and it was found 
that in a four cylinder motor thus arranged and subjected to 
a blast of air from a fan located in front, cylinder No. 3 
reaches the highest temperature. 

Blower Cooling—These different problems have been satisfac¬ 
torily solved, however, by a system known as blower cooling. 
The cylinders are formed with vertical flanges or cooling pins 
of relatively small height, and these are surrounded by an 
aluminum or sheet metal air jacket. The jackets of the indi¬ 
vidual cylinders connect on top to an air pipe leading to the 
outlet of a centrifugal fan or blower. In the Franklin motors 


r' •>> j 



Fig. 258. 






AIR COOLING. 


403 


the flywheel is made in the form of a centrifugal fan, while 
in the Kelly motor, illustrated in Fig. 259, a blower is located 
at the forward end of the motor. In this way a strong blast 
of cool air is directed against the heads of the cylinders 
which are heated the most, and the cooling effect is the same 
for all of the cylinders and all around each cylinder. This 
system resembles water cooling in many respects. There is 
this difference, however, that the air need not be stored, but 



is always and everywhere available in unlimited quantity. 
Leakage troubles and troubles due to freezing of the cooling 
water in winter are entirely eliminated. A further difference 
is that whereas with water cooling the cylinder wall tempera¬ 
ture is limited to 212 0 , with air cooling it will generally be 
considerably higher than this, which has a favorable effect 
on the fuel economy. The temperature of the cylinder walls, 
of course, will depend upon the capacity of the blower. 

The Blower—Suppose that a 25 horse power motor is to 
be cooled by the blower system. We may assume that the 
fuel consumption is one pound of gasoline per horse power- 
hour (taking a rather high figure for the sake of safety), and 
that 50 per cent, of the fuel energy is carried off by the cool¬ 
ing air. What must be the delivery of the blower at the full 


































































































































404 


AIR COOLING. 


load speed of the motor? The heat energy to be carried off 
by the cooling air is evidently 

2 - ^ ^~ J 9 ,000 = 4,000 B. T. U. {apfr.) per minute. 

2 X 60 

Now, the specific heat of air at constant pressure is 0.24. 
Allowing for a rise of ioo° Fahr. in the temperature of the air, 
the amount to be moved is 



Fig. 260.—Centrifugal Fan for Air Cooled Motor. 


—- = 167 pounds. 

100 X 0.24 

Since atmospheric air weighs 0.08 pound per cubic foot, the 
amount to be moved is 

- 7 - — 2,100 cubic feet per minute. 

0.08 































AIR COOLING. 


405 


The pressure against which this air must be moved depends 
upon the dimensions of the jackets. 

A centrifugal fan or blower consists of two parts, viz., the 
casing, whose peripheral surface is of spiral shape, and the 
runner. The latter is a revolving wheel with numerous blades 
or vanes on its circumference which cause the air to re¬ 
volve with the runner. The air enters through the open side 
of the casing (see Figs. 259 and 260), has a rotary motion 
imparted to it by the blades, is forced outward against the 
casing spiral by its centrifugal force, and because it strikes 
the spiral at an angle, is forced in the direction of the outlet. 
Centrifugal fans are high speed machines, and when gear- 
driven from the engine run at 2 to 2^4 times crankshaft 
speed. They may be driven either by belt or silent chain. 
When gears or chains are used it is necessary to drive through 
a friction clutch, owing to the great inertia of the runner, 
depending upon its high speed, which would throw heavy 


1 -V, 


j- 

--Ti 



Fig. 261.—Exhaust Valve Housing. 

strains upon the drive and the runner itself if the motor were 
abruptly speeded up or loaded down. 

Air Cooled Exhaust Valves—The exhaust valves in an air 
cooled motor—unless auxiliary exhaust ports are used—are 
subjected to unusually severe service, owing to the high tem¬ 
perature at which they necessarily work. For this reason 
the use of valve cages inserted into pockets in the cylinder 





















































406 


AIR COOLING. 


head has been found more or less impractical. It is a better 
plan to make the valve seats in the shape of an L fitting, one 
branch of which is bolted to the cylinder head and the other 
connected to the exhaust pipe, as shown in Fig. 261. These 
valve casings can be cast with cooling ribs so as to further 
increase the cooling effect. 


CHAPTER XX. 


THE TWO STROKE CYCLE MOTOR. 

The two cycle motor has been and still is used very extensively 
for marine work, but its use on automobiles has always been lim¬ 
ited. At present there' are three somewhat prominent firms in the 
United States using this motor on automobiles, besides a few less 
well known ones. 

The characteristic feature of the two cycle motor is that there 
is an explosion in each cylinder during every down stroke of the 
piston, instead of during every second down stroke, as in the four 
cycle motor. One other feature which distinguishes the two cycle 
motor from the four cycle is that the combustible charge must be 
pre-compressed in order to get it into the working cylinder. This 
is generally done in the crank case, which is made gas-tight and 
acts as a pump; but it is sometimes done by means of a pump 
formed by extensions of the working cylinder and piston or by 
an entirely separate pump cylinder. As originally adapted from 
marine practice the two cycle motor was of extreme simplicity, 
but it possessed certain defects, and in order to overcome these 
it was found necessary to add parts, thus complicating the motor 
to a certain extent. 

The Two Stroke Cycle—The successive operations in the 
cylinder and crank case of a two cycle motor are as follows: 
Referring to Fig. 262, while the piston is moving up in the cylin¬ 
der it creates a partial vacuum in the crank chamber, and when 
it reaches a certain point in its up-stroke it uncovers a port A in 
the wall of the cylinder, through which combustible charge from 
the carburetor is drawn into the crank chamber. During the fol¬ 
lowing down stroke, after the port A has been covered by the 
piston, this combustible charge is slightly compressed in the crank 
chamber, the usual compression pressure being between 6 and 8 
pounds per square inch, gauge. As the piston approaches the 
bottom end of its stroke it uncovers the transfer port B in the 
cylinder wall (see Fig. 263), and the combustible charge under 
pressure in the crank chamber flows through the transfer passage 


407 




408 


THE TWO STROKE CYCLE MOTOR. 


into the cylinder. Shortly after the piston starts on its up¬ 
stroke it closes the transfer port, and during the remainder of 
the up-stroke the charge is compressed in the cylinder. When 
the piston reaches the top end of the stroke the charge in the 
cylinder is ignited and the piston is forced down. As the piston 
approaches the lower end of its stroke it uncovers the exhaust 
port C in the cylinder wall opposite the transfer port B (see 
Fig. 264), and the burnt gases, which are still under considerable 
pressure, escape through this port. It should be pointed out that 


Fig. 262.—Admission to 
Crank Case. 




Fig. 263.—Transference of 
Charge. 


the exhaust port opens slightly earlier than the transfer port, and 
a considerable portion of the burnt gases escapes before any new 
charge enters through the transfer port. The rest of the burnt 
gases are swept out by the incoming fresh charge. It is, of 
course, impossible to insure that all of the burnt gases be swept 
out and that none of the fresh charge escape; but in order to 
prevent as far as possible the mingling of the fresh charge with 























































































THE TWO STROKE CYCLE MOTOR. 


409 


the spent gases a deflector D is formed on the piston head oppo¬ 
site the transfer port which deflects the incoming current of 
fresh gas upwardly (Fig. 263). 

The evacuation of the cylinder and the transference of a new 
charge from the crank chamber to the cylinder, therefore, go on 
simultaneously; and, besides, the periods of admission and ex¬ 
haust, measured in degrees of crank motion, are much shorter 
than in the four cycle motor. While in a four cycle motor the 
admission extends over a period corresponding to about 200 de 
grees of crank motion, and the exhaust to a period correspond¬ 


ing to 230 degrees of crank 
motion, in a two cycle motor the 
inlet extends over less than 90 
degrees of crank motion, and 
the exhaust over less than 120 
degrees of crank motion. 

Crank Case Inlets—In the 
older designs of two cycle mo¬ 
tors the charge from the car¬ 
buretor was admitted to the 
crank chamber through a suction 
operated poppet valve. This type 
is known as the two port motor, 
and is extensively used for 
marine work. It has, however, 
never found any application for 
automobile purposes, owing to 
the fact, it is claimed, that it does 
not lend itself to operation at 
high speed. Since the degree of 
vacuum in the crank chamber is 
always quite low, the valve ac¬ 
tuating force is very small, and 
it would be expected that at high 
speeds the inertia of a suction 
actuated valve would be a dis¬ 
turbing factor in the operation 
of the motor. 



The type of two cycle motor illustrated in Figs. 262-264 is 
known as the three port type. In it a port is cast in the cylin¬ 
der wall at such a point that it will be uncovered by the lower 
end of the piston when the latter approaches the top end of its 
stroke, and since at that time there is a partial vacuum in the 
crank chamber, combustible charge from the carburetor is drawn 








































410 


THE TWO STROKE CYCLE MOTOR. 


through the port into the crank chamber. Such a port opens 
and closes positively, but it remains open only a comparatively 
short length of time. The suction on the carburetor starts and 
stops abruptly, which condition calls for a special design of 
carburetor. 

A form of crank chamber inlet which combines positive op¬ 
eration with gradual opening and closing and an extended 
period of opening is shown in Fig. 265. In this design a flat 
plate A with an inlet port cut in it is bolted to an opening in 
the crank case. A flat valve disc B with a valve port adapted to 
register with the port in the plate A is placed against the plate 
from the inside and in driving connection with a shaft which is 
driven at crankshaft speed. The driving means consists of a 
crossbar C pinned to the shaft and provided with two driving 
pins D D and two spring pressed plungers E E for driving the 



disc and keeping it in contact with the valve plate, respectively. 
With this type of inlet valve, charge may be admitted to the 
crank chamber during 135 degrees of crank motion, that is, from 
the time the transfer port closes to the end of the up-stroke, 
or even a little longer. The disc valve may be placed concentric 
with the crank journal and driven directly from the crank arm, 
only its rubbing speed then becomes rather high. 

A somewhat similar port arrangement consists in forming ports 
in one of the crankshaft journals and in its bearings. (See Fig. 
272.) These ports can be timed the same as the disc valve just 
described. 

Crank Chamber Design—The crank chamber of a two cycle 
motor differs from that of a four cycle in many respects. In the 
first place, it must be made as small as possible, or the necessary 
















THE TWO STROKE CYCLE MOTOR. 


411 


crank chamber compression will not be obtained. If we assume 
that the crank chamber fills to a pressure of 14 pounds per square 
inch absolute and we wish a precompression of 7 pounds per 
square inch gauge, or 21.7 pounds absolute, then the required 
crank chamber volume may be calculated as follows: With this 
low compression iatio it is permissible to assume that the pressure 
varies inversely as the volume. Compression begins when the 
inlet port closes and ends when the transfer port opens. _ We 
will assume that the motor is of 4 inches bore and 4^ inches 
ttroke, that it is of the three port type, that both the inlet and 
transfer ports are Yz inch wide and that their outer edges are even 
with the limit of piston travel. Compression then begins when the 
piston has traveled l /z inch on its down stroke and ends when it 



Fig. 266.— Two Stroke Crank Chamber. 


has traveled 4 inches, and has still l /z inch niore to travel. Now, 
let us call the initial volume of the charge Vi and the final 
volume Vi. Then the reduction in the volume of the charge 
during compression is equal to the volume of a cylinder 4 inches 
in diameter and 3 Y* inches long, which is 

4 ^ X 3.14 X 3.5 = 43-96 cubic inches. 

4 

Vi 2 1 .7 

. Vi 14 
V\ — Vi = 43.96 


Consequently 



















412 


THE TWO STROKE CYCLE MOTOR. 

V\ — i. 55 Vt 
0-55 V* = 43-96 
Vt = 80 cubic inches. 

That is, the capacity of the crank chamber at the moment the 
transfer port opens, when the piston is J/2 inch from the bottom 
end of the stroke, must be 80 cubic inches. 

In order to get the crank chamber down to this volume, various 
expedients are resorted to. In the first place, the crank chamber 
wall is made to come as closely as possible to the form described 
by the crank in its revolution. Then the connecting rod is made 
as short as possible, say, 1.8 times the stroke, and if this does not 
bring the volume down sufficiently a crankshaft with discs instead 
of crank arms may be used. 

A manograph diagram from the crank case of a two cycle 
motor, taken by Albert L. Clough, is shown in Fig. 267. At the 
beginning of the up-stroke, there is a slight over-pressure in 
the crank case, and this increases during the first part of the 



stroke, as long as the transfer port remains open, which may be 
due to the heating of the charge as it comes in contact with the 
cylinder walls, or to a return wave following the surging of the 
charge into the cylinder. Shortly after the transfer port is closed 
a vacuum begins to form in the crank chamber, increasing up to 
the point where the inlet port opens. The crank chamber then 
fills rapidly and at the end of the stroke is filled nearly to 
atmospheric pressure. During the return stroke the charge is 
compressed in the crank chamber and attains a maximum pressure 
of about 8.5 pounds per square inch at the point where the 
transfer port begins to open. The pressure thereafter rapidly 
drops to near atmospheric. 

Since the crank chamber acts as a pump it must be made and 
maintained air-tight. In the three port type of engine it is neces¬ 
sary to place one packing ring on the piston near the lower end, 
so that the charge under pressure in the crank chamber will not 
blow back by the piston and out through the inlet port. There is 
also chance for leakage through the crankshaft bearings. In order 







THE TWO STROKE CYCLE MOTOR. 


413 


to minimize this leakage the bearings must be made of liberal size, 
and particularly of good length, so they will wear very slowly. 
Grease lubrication of the bearings tends to keep the charge in, but 
is not very suitable for automobile work. Force feed of oil to the 
bearings is the next best thing. Some manufacturers provide the 
end bearings with packing on the outer ends. 

In a multicylinder two cycle motor the crank chamber of each 
cylinder must be separate, in order that it may act as a charge 
pump. 

Transfer Passage and Deflector—From the crank chamber 
the charge must be transferred to the cylinder while the piston 
is at and near the lower end of its stroke. Sometimes the trans¬ 
fer passage extends down the side of the cylinder and through the 
top portion of the crank chamber, as seen in Figs. 262-264. How¬ 
ever, certain designers prefer to shorten the passage by terminat¬ 
ing it below in a port in the cylinder wall with which registers a 
port in the piston wall when the piston is at the bottom end of 
its stroke. (See Fig. 268.) 

The cross section of the transfer passage should be at least 
equal to that of the transfer port. All of the three ports extend 
over 80 to 85 degrees of the cylinder circumference, but their ef¬ 
fective width is reduced by a bridge at the middle about inch 
wide. This bridge is necessary in order to prevent the piston 
rings from springing out and being caught by the edges of the 
port. The bridge in the exhaust port naturally becomes very hot 
and in cylinders of considerable 
size it is customary to make it 
hollow and cool it by either 
water or air circulation. 

It is a good plan to insert a 
wire screen in the transfer pas¬ 
sage or place it against the open¬ 
ing of same into the crank cham¬ 
ber, so as to prevent explosions 
in the crank chamber. When 
the charge admitted to the cyl¬ 
inder is low in gasoline it burns 
very slowly and is still burning 
when the transfer port is opened. 

The new charge arriving from 
the crank chamber is then easily 
ignited, resulting in an explosion in the crank chamber whereby 
the three or four following explosions are missed. The same 
thing occurs when the spark is set late. This can be prevented by 



Fig. 268.—Transfer Port in 
Piston Wall. 















































THE TWO STROKE CYCLE MOTOR. 


414 

placing a wire gauze screen of sufficiently fine mesh across the 
transfer passage, which, on the familiar principle of the miner’s 
lamp, will prevent the flame from passing through it. The screen, 
of course, must be sufficiently rigidly secured so as to withstand 
the forces acting upon it, and it must be of such large area as not 
to throttle the charge. 

In order to prevent the fresh charge from mingling with the 
spent gases a deflector plate is usually formed on the piston head. 
The form of this deflector differs considerably in different engines. 
Its height is generally limited by the height of the compression 
space. Sometimes, in order to increase the permissible height, the 
piston head is made of irregular shape, as shown at A in Fig. 269, 



Fig. 269. —Deflectors. 


the deflector forming part of the piston head wall. The distance 
of the deflector plate from the cylinder wall is generally slightly 
greater than the height of the transfer port. The central part 
of the deflector generally forms an arc of a circle concentric with 
the cylinder, while the ends swing around toward the cylinder 
wall, as shown at B in Fig. 269. It is well to provide a liberal 
fillet where the deflector joins the piston, so no eddies will be 
produced in the corner and interfere with the rapid flow of the 
charge. 

Port Sizes—The operation of a two cycle motor depends 
very largely upon the dimensions and location of the different 



















THE TWO STROKE CYCLE MOTOR. 


415 


ports. The ports must be located very accurately in the casting, 
for the reason that once the motor is completed it is impossible 
to make any adjustments, as is possible in a four cycle motor, for 
instance. The dependence of the port width and height upon the 
bore and stroke of the motor is a matter of great importance. 

If we take two motors with cylinders of different dimensions, 
intended for the same speed of revolution, then the capacities for 
passing gas of corresponding ports should evidently be propor¬ 
tional to the cylinder volumes. However, as already found in con¬ 
nection with four cycle motors, if the stroke varies the speed of 
revolution will vary. It is logical to assume that, the same as in a 
four cycle motor, the speed of rotation will vary inversely as the 
square root of the length of stroke. The port capacity must, of 
course, increase directly" with the speed of rotation. Then, de¬ 
noting the port capacity by Q, we may write 

Q~ J X |/7 ~ iV 7 .(99) 

An analysis of a number of successful two cycle motors shows 
that the average value of the transfer port capacity is 

Qt = 2 b 2 +/ i square inch-degrees .(ioo) 

and the average value of the exhaust port capacity 

Q e = 3.3 5 2 < v/ l square inch-degrees ...( 101 ) 

Calculation of Port Capacity—Referring to Fig. 270, it will 
be seen that the distance 0 by which the port is opened may be 
expressed by the equation 

o — h — (/ — x) = h + x — l. 

Inserting the value of x found in equation (20) — 

o = h - 4 - —_ - cos 9 -f- sin 2 9 — l — h -— — cos 9 -f-— sin 2 9. 

228 n 228 n 


In Fig. 271 is shown a co-ordinate 
diagram in which the abscissas repre¬ 
sent crank angles and the ordinates 
length of port opening. Let d 9 repre¬ 
sent any small increment of the crank 
angle and 0 the corresponding length 
‘of opening of the port. Then the prod¬ 
uct 0 d 9 represents the area of the 
narrow strip of width d 9 and height 0. 
The total area enclosed by the port 
opening curve and the horizontal axis 
may be called the port opening 


^p\\\\\\\\\\\\\\\\\^ 




Fig. 270. 

integral, and be denoted by /. Inserting the value of 0 in the 























416 THE TWO STROKE CYCLE MOTOR. 


expression for the area of the narrow strip, we have 

o d 0 = h d 6 — — d 0 - - — cos 6 d 6 - 1- — sin 2 6 d 6 . 

2 2 8 n 

In order to find the area of the entire diagram we have to inte¬ 
grate this expression between the limits 0i and # 2 : 


f 


*2 


h d Q _— d 6 — — cos 6 d 6 -)—sin s 8 d 0 = 

2 2 8k 


^/i — L\ /# 2 — 6^ — — ^sin @ 2 '— sin 

4- —-— (# 2 — 0i — sin d 2 cos 0 ., + sin 6 X cos 6 X ).(102) 

16 n \ ‘ / 

In applying this equation the angles #1 and 0 2 must be expressed 

in radians and the result will be in inch-radians. To change it 

to inch-degrees it is only necessary to multiply by=^-. However, 

6.28 

equation (102) can be materially simplified. The angles and #2, 
of course, represent the angles of crank motion from the begin¬ 



ning of the down stroke to the moments of port opening and 
closing, respectively. Now let a represent the angle of crank mo¬ 
tion corresponding to the period from the beginning of port open¬ 
ing to full opening or the end of the down stroke. Then 

0 2 — #i = 2a 

Sin #2 — sin #1 = — 2 sin a 

(remembering that #2 is in the third quadrant and #1 in the second) 
— sin #2 cos 02 + sin #1 cos #1 = — 2 sin a cos a. 
Substituting in equation (102) we have 


/= (*~j) 2a + 7 ( 2Sina ) + l6 n { 2a ~ 


2 sin a cos 




(2 h — /) a -f- /sin a + — (a — sin a cos a). 

8 n 


This equation, as already stated, gives the opening integral in 
inch-radians. In inch-degrees the opening integral is 


■_ 360 - * 1 











THE TWO STROKE CYCLE MOTOR. 


417 


But in order to avoid the necessity of first transforming the 
angular measurements from degrees into radians we may trans¬ 
form this equation as follows: 


j— 360 
6.28 


360 

6.28 


l sin a 


(■ 


( 


l sin a — — sin a cos a 

8 n 1 



a 



8 n 


ain.-degs 


(103) 


In applying this equation the value of « in degrees must be in¬ 
serted. 

Equation (103) involves two dependent variables, viz., h 
and a. In order to make use of it, it is, therefore, necessary 
to first assume a value for h and then find the corresponding 
value of « by mtans of equation (49) — 

Cos 6 = 1/ 4 rz 2 —|- 4 « —j— / — 8 n -j 2n - 

The angle a is the supplement of angle 0 , or 180 degrees 
— 0 . However, since both of these equations are rather un¬ 
handy for use, the values of a and I for a wide range of 
strokes and port heights have been calculated and are given 
in Tables VII and VIII respectively. In calculating these 
tables the ratio n of connecting rod length to length of stroke 
was assumed to be 1.8. The values given in the tables, how¬ 
ever, are not changed much by a slight change in this ratio. 
The following example illustrates the use of these tables and 
of equations (100) and (101). 

Suppose that it is required to find the proper dimensions 
for the transfer and exhaust ports of a 4^x5 inch two stroke 
engine. From equation (100) we find that the transfer port 
capacity should be 

2X 4 - 5 2 J 5 =90.6 square inch-degrees , 


and from equation (101) that the capacity of the exhaust 
port should be 

3-3X4-5 2 ' / 7 = 149-5 square inch-degrees. 

Now, in order to get the proper timing the transfer opening 
should extend over a little less than 90 degrees of crank mo¬ 
tion (say a= 43 degrees), and the exhaust opening over 
about 115 degrees of crank motion (<1 = 57.5 degrees). By 
referring to Table VII we see that with a stroke of 5 inches 
a port height of V2 inch gives for a a value of 42^ degrees 
and a port height of 7 /g inch, a value of 56 degrees 26 minutes. 
We will adopt these values. By now referring to Table VIII 
we see that the port opening integral for 5 inch stroke and 






TABLE VII—PORT 


4J8 


THE TWO STROKE CYCLE MOTOR 


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THE TWO STROKE CYCLE MOTOR. 


419 


l / 2 inch height of port is 28.5 inch-degrees, and for 5 inch 
stroke and % inch height of port, 65.1 inch-degrees. By di¬ 
viding the necessary port capacity found above by these 
values for the port integral, we get the required circumferen¬ 
tial width of the ports. For the transfer port 

zut = 9 °• 6 — 3. j8 inches, 

28.5 

and for the exhaust port 

zut — T - — 5 = 2.3 inches. 

65.1 

As used for marine purposes the two stroke motor is pro¬ 
vided with ports much smaller than those here calculated, but 
marine two cycle motors will run only at comparatively low 
speed and furnish only a small output in proportion to their 
cylinder dimensions. To adapt this motor to high speed op¬ 
eration for use on automobiles the ports had to be enlarged. 
This, of course, involved some disadvantages. The effective 
stroke is reduced by the height of the exhaust port, and the 
greater this height the shorter the effective stroke. More¬ 
over, when the ports are very high, it is impossible to pre¬ 
vent some of the fresh charge from escaping through the 


TABLE VIII—PORT OPENING INTEGRALS. 


Stroke. 

Port Height. 

1 % inches.... 

6 

Ins. 

5H 

Ins. 

5'A 

Ins. 

5% 

Ins. 

5 4H 

Ins. Ins. 

4'A 

Ins. 

4'A 

Ins. 

4 

Ins. 

101.5 

.... 

. . . . 


JL/ v. g I L vo> 

• • • • 

.... 

• • • «* 

1 3-16 inches.. 

93.7 

95.9 

98 

.... 

.... 

.... 

.... 

.... 

.... 

1% inches.... 

86.4 

88.3 

90.8 

92.4 

94.5 

.... 

.... 

.... 

. . . - 

11-16 inches.. 

80.3 

81.7 

83.4 

85.3 

87.5 

90 

92.7 

.... 

..... 

1 inch. 

73.4 

75 

76.7 

78.4 

80.4 

82.8 

85.6 

87.5 

89.9* 

15-16 inch. . .. 

66.8 

68.1 

69.6 

71.3 

73.2 

75.1 

77.1 

79.2 

81.6- 

% inch. 

60 

61.5 

63 

64.3 

65.1 

67.5 

69 

71.1 

73.4- 

13-16 inch.... 

53.9 

55.3 

56.8 

57.8 

59.3 

61.1 

62.4 

64 

65.9 

A inch. 

47.7 

49.2 

50.3 

51.4 

53.2 

54.2 

55.4 

56.8 

58. S 

11-16 inch.... 

41.9 

43.2 

44.3 

45 . 4 

46.4 

48.1 

49.2 

50.3 

51.4 

$4 inch . 

36.3 

37.6 

38.5 

39.3 

40.3 

41.9 

42.4 

43.6 

44.6^ 

9-16 inch . 

32.1 

32.1 

32.9 

33.6 

34.5 

35.4 

36.5 

37.5 

38.4 

y 2 inch . 


26.9 

27.2 

27.8 

28.5 

29.3 

30.4 

31.3 

32.3 















420 


THE TWO STROKE CYCLE MOTOR. 


exhaust port, whereby the fuel economy will be reduced. 
There is, therefore, a limit to the practical size of the ports, 
and this limit is supposed to be given by equations (ioo) 
and (ioi). 



tional to the cylinder volume, and in the calculation of the 
port capacity it was assumed that this capacity for passing 

















































































































































THE TWO STROKE CYCLE MOTOR. 


421 


charge is proportional to the area of the port. This assump¬ 
tion is probably not absolutely correct. According to works 
on hydraulics, the discharge from an orifice in the wall of a 
vessel is proportional to the area, while the discharge through 
a pipe of considerable length is proportional to the 2.5th 
power of the diameter, hence, varies slightly faster than the 
cross sectional area of the pipe. The transfer and exhaust 
passages present cases intermediate between those of orifices 
and pipes, and it might, therefore, appear that the capacity 
for passing charge would increase slightly faster than the 
area. However, when it is considered that in a motor the 
flow is periodically started and stopped, and that, conse¬ 
quently, the quantity of charge passed depends to some ex¬ 
tent upon the inertia of the gas, and, besides, that the length 
of the passages generally increases with the cylinder dimen¬ 
sions, it will be seen that the assumption that the quantity 
of gas passed in a given time is proportional to the port area 
must be very near the truth. Ports dimensioned according 
to equations (100) and (101) give a constant gas velocity in 
cylinders of all dimensions, if the degree of filling or the 
volumetric efficiency remains constant. 

It should be pointed out that up to now considerable diffi¬ 
culty has been experienced by manufacturers who have suc¬ 
ceeded in properly working out an engine of one particular size 
in determining the proper sizes of ports for an engine of different 
cylinder dimensions, especially if the ratio between the length 
of stroke and piston head area was different. For this reason 
most manufacturers of three port, two stroke automobile 
motors have adhered closely to two sizes of cylinders, viz., 
4x4]^ inch and 4L> X 5 inch. 

The ports should be finished by means of a milling cutter 
of small diameter, which is run right through them, so as to 
get them of the correct width and at the correct distance 
from the bottom face of the cylinder casting. 

There is one disadvantage in the usual arrangement of 
the ports, as illustrated in Figs. 262-264, and that is that the 
greater heating of the cylinder wall on the exhaust port 
side tends to distort the cylinder. Some manufacturers re¬ 
cently have sought to overcome this defect by providing two 
oppositely located transfer ports and two exhaust ports in¬ 
termediate between the transfer ports. While this overcomes 
the trouble due to unequal expansion, it reduces the distance 
between transfer and exhaust ports, and therefore increases 
the liability of leakage between these ports. (See Fig. 272.) 


422 


THE TWO STROKE CYCLE MOTOR. 


Carburetors and Throttling —Owing to the fact that the 
suction in a three port, two stroke motor, lasts less than half 
as long as in a four stroke motor, it is proportionately 
stronger. For this reason a carburetor adjusted for a four 
stroke motor generally does not give very satisfactory re¬ 
sults on a three port, two stroke motor. It has been found 
that if a carburetor of the supplementary air valve type is to be 
used on the latter type of motor, the primary air inlet around 
the fuel jet must be made very small, or there is likelihood 
of flooding at low speed. 

The throttle valve can be placed either in the transfer port 
or in the inlet pipe at the carburetor. From a manufactur¬ 
ing standpoint the latter location is preferable, and if the 
carburetor is properly designed and adjusted it is equally sat¬ 
isfactory. One important difference between the two stroke 
and four stroke motors is that when the two stroke motor is 
throttled its compression is not reduced. The combustion 
chamber at the beginning of the compression stroke is always 
filled substantially to atmospheric pressure, and since less 
new charge is allowed to enter, more spent gases remain. 
This, of course, results in a dilution of the charge and ren¬ 
ders ignition more difficult. The location of the ignition 
terminals is for this reason of greatest importance. If the 
spark plug is placed in the side wall of the cylinder directly 
above the transfer port it is most certain to be surrounded 
by fresh charge under all conditions, but unfortunately its 
points are also directly exposed to any oil globules carried 
along by the incoming charge, for which reason some makers 
prefer to place the plug in the centre of the cylinder head. 

Four Cycling —When a two stroke motor of ordinary con¬ 
struction is throttled down much it usually begins to miss, 
which is due to excessive dilution of the charge. Sometimes 
throttling gives rise to a phenomenon known as “four cy¬ 
cling.” That is, the engine will miss every second charge. 
This phenomenon is not difficult to explain. Evidently, when 
a charge is missed it contains too large a proportion of dead 
gas. During the next transfer period a new charge enters, 
which, instead of being mixed with dead gas, is mixed with 
a diluted charge, and this new charge, therefore, is not di¬ 
luted so much that it is not inflammable. The following 
charge, however, is again mixed with dead gas, and therefore 
fails to explode. 

Lubrication —In two stroke motors with crank case com¬ 
pression it is impractical to employ splash lubrication, since 


THE TWO STROKE CYCLE MOTOR. 


423 


a great deal of the lubricant would then pass with the charge 
into the combustion chamber and would cause a smoky ex¬ 
haust, besides waste of oil. Many two cycle motor manu¬ 
facturers therefore lubricate their cylinders by either mix¬ 
ing lubricant with the gasoline or feeding it to the inlet pipe 
just beyond the carburetor. When mixing the oil with the 
gasoline about one quart of oil is used for every five gallons 
of fuel. The crankshaft journals may be lubricated by 
means of a mechanical oiler, and arrangements may be made 
for conducting the overflow from these bearings to the 
crank pin bearings. Some manufacturers provide a wick oil 
feed for the crank pin bearing, the wick rubbing against the 
crank chamber wall and collecting the oil which precipitates 
on the wall from the fuel charge. One point to be observed 
in engines with crank case compression is that the oil 
grooves in the crankshaft bearings must not be brought top 
close to the outer ends of the bearings, or the oil will be 
blown out through the bearings by the compression. These 
bearings are generally babbitt lined, the babbitt being cast 
directly into the bearing hubs, and the bearings then reamed. 

Differential Pistons—In order to avoid the difficulties in¬ 
herent in crank case precompression, differential pistons have 
lately been used to quite an extent. It may be pointed out 
at once that this construction is applicable only in the case 
of engines with an even number of cylinders. By referring 
to Fig. 273, which shows an engine of this type, it will be 
seen that the cylinder is made with a small bore at the top 
end and a larger bore at the lower end. The piston is also 
made with a double diameter, the small diameter portion on 
top fitting the small bore of the cylinder and the larger diam¬ 
eter, lower portion fitting the bigger bore. The regular work¬ 
ing cycle takes place in the upper end of the cylinder above 
the piston head. An annular space is formed between that 
portion of the cylinder wall which is bored to the larger 
diameter, the small diameter portion of the piston and the 
shoulders or offsets in the piston and cylinder walls, re¬ 
spectively. As the piston moves down the combustible charge 
is drawn into this annular space, either through a suction 
operated poppet valve or a port in the cylinder or piston 
wall. If the port is in the piston wall the charge must natur¬ 
ally pass through the crank chamber. Upon the return 
stroke of the piston the valve or port closes and the charge 
is compressed in the annular space to a relatively high pres¬ 
sure. The ports would be so proportioned that at the mo- 


424 


THE TWO STROKE CYCLE MOTOR. 


ment the transfer port is opened the distance of the offset 
in the piston from the offset in the cylinder is equal to 
about one-quarter the length of stroke. When the piston 
approaches the end of its up-stroke the annular space is 
placed in communication with the upper end of another cyl¬ 
inder in which the piston is at that moment approaching the 
end of its down-stroke, and which is, therefore, ready to 
receive a new charge. This method of construction is par- 



Fig. 273.— Cote Differential Piston, Two Stroke Motor. 


ticularly applicable to double cylinder motors. In these 
the two cranks are set at 180 degrees, and one piston is at 
the top end of the stroke while the other is at the bot¬ 
tom end. The port in the wall of the annular chamber of one 
cylinder communicates through a passage with the working 

















































































THE TWO STROKE CYCLE MOTOR. 


425 


end of the other cylinder, and when the transfer port of the 
latter cylinder is uncovered by its piston the charge from the 
annular compression space is transferred to the working end 
of the other cylinder. One method of thus cross-connecting 
the compression and combustion chambers of two cylinders 
is illustrated in Fig. 274. 

The displacement in the annular space would generally be 
made a little greater than the displacement in the working 
end of the cylinder, with the object of getting more charge 
into the cylinder during each revolution. For instance, with 
a bore of 4 inches for the working end of the cylinder, the 
bore of the annular space would be made 6 inches, which 
gives an increased displacement of about 25 per cent. 

One objection to the differential piston construction is that 
with a normal length of stroke it gives a rather high engine. 




[ f 




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Section one -d 

Fig. 274— Transfer Passage of Cote Motor. 


which calls for a high, unsightly bonnet and raises the centre 
of gravity. Another disadvantageous feature is that the re¬ 
ciprocating parts are of necessity quite heavy and make high 
speeds impractical. With these pistons the problem also 
arises as to where to place the piston pin. If it is placed 
about midway of the length of the piston (which would be 
the best location from the standpoint of pressure distribu¬ 
tion), there might be leakage from the annular chamber to 


































































































































426 


THE TWO STROKE CYCLE MOTOR. 


the crank chamber around the ends of the piston pin. For 
this reason some manufacturers place the piston pin in the 
top bearing portion of the piston, others in the bottom bear¬ 
ing portion, and still others just below the top ring covered 
portion, so that it is uncovered only when the piston is near 
the bottom of its stroke, when the pressure in the aDnular 
space is near atmospheric. 

Four Cylinder, Two Stroke Motors—In earlier years two 
stroke motors for automobiles were built only in single 
and double cylinder models, but owing to the general pref¬ 
erence for four cylinder construction, quite a number of four 
•cylinder, two stroke motors have recently been built. In these 
it is customary to place the two cranks for each end pair of 
•cylinders at 180 degrees, and the two sets of cranks at 90 
degrees with each other. This causes the explosions to 
occur at equal intervals, but does not result in an inherent 
balance of the crankshaft, and balancing weights should pref¬ 
erably be used. Two ..carburetors are generally used with 
this type of motor, one with each end pair of cylinders. Such a 
motor naturally produces a very uniform turning moment. 
Since the exhausts in a four cylinder motor overlap consider¬ 
ably, a separate exhaust pipe has to be provided for each 
pair of cylinders. 

Distributor Valve Motor—In some of the more recent two 
•cycle automobile motors rotary valves are used in con¬ 
junction with the double diameter cylinders. One of 
the earliest motors of this type is the Legros, made in 
France, which is illustrated in Fig. 275. In this engine there 
is a stationary piston A inside the regular working piston B, 
and the charge is compressed between the stationary and 
movable piston B. The stationary piston is cast with a pas¬ 
sage C leading to the distributing valve chamber. This 
valve D establishes communication either between the car¬ 
buretor and the passage C in the fixed piston or between the 
compression chamber E and the passage C. Suppose the 
movable piston to start from the top end of the stroke under 
the force of an explosion. As it approaches the bottom end 
of the stroke the spent gases will be discharged through the 
exhaust port, and a moment later a new charge will enter 
the working cylinder from the compression chamber E 
through the transfer port. During the following up-stroke 
of the piston the new charge will be compressed in the 
working cylinder. At the same time the space between the 
•movable and stationary pistons is in communication with the 


THE TWO STROKE CYCLE MOTOR. 


427 


carburetor through the rotary valve D, and a charge is drawn 
into this space. As the piston reaches the upper end of its 
stroke the rotary valve closes the passage to the carburetor 
and instead places the passage in the stationary piston in 
communication with the compression chamber E in which 
the charge is compressed during the down-stroke of the 



movable piston. From this compression chamber, as already 
explained, the charge passes into the working cylinder upon 
the opening of the transfer port. In this construction, too, 
the connection of the connecting rod with the piston does 
not seem to be entirely satisfactory, as it is at the extreme 
lower end of the piston, so that the side thrust pressure is 
very unevenly distributed. 














































































428 


THE TWO STROKE CYCLE MOTOR. 


In the Elmore High Duty engine, which is best known in 
this country, a rotary valve is used which places the annu¬ 
lar compression space alternately in communication with 
the carburetor and the transfer port of an adjacent cylinder. 
Referring to Fig. 276, the incoming gas is drawn into the 
annular chamber D during the entire downward stroke of the 
piston. The gas passes from the carburetor through the 
manifold A into the distributor valve E, thence through the 
distributor port and inlet passage C, and fills the annular 
chamber D around the piston below the firing chamber. 
During the following up stroke of the piston the gas in 
chamber D is compressed and is forced through the transfer 



Fig. 276.— Elmore High Duty, Two Stroke Motor. 


passage into the distributor valve, which has now changed 
its position to admit the new gas; thence it flows through 
transfer port F into the combustion chamber of the adjacent 
cylinder. 

Horse Power Output—As in the case of four stroke mo¬ 
tors, it is impossible to give a satisfactory formula for the 
output of two stroke motors. As a rule, the two stroke 
motor gives its maximum output at a lower piston speed 
than the four stroke motor, and its mean effective pressure 
is also somewhat less. But since there are relatively twice 
the number of explosions in a two stroke motor, the latter 
may give a greater power output for the same cylinder 
dimensions. A two cylinder, three port 4x5 inch engine 





































































































THE TWO STROKE CYCLE MOTOR. 


429 


tested at the University of Michigan gave a maximum output 
of 13.5 horse power at 930 r. p. m. E. W. Roberts, in a 
paper read before the mechanical branch of the Association 
of Licensed Automobile Manufacturers at New York in 
July, 1908, stated that he had obtained 20 horse power from 
a two cylinder motor of these dimensions. From a single 
cylinder 33 4 x 3/4 three port engine Dr. Watson obtained a 
maximum of 434 horse power at 1,500 r. p. m. The power 
of two French motors is given in the following paragraph. 

Tests of Two Stroke Motors —A contest for two stroke 
motors was conducted by the Automobile Club of France 
during October and November, 1907. Seven different mo¬ 
tors were submitted for the tests, but only two met all of 
the conditions. The test comprised a six hour run at full 
load, a three hour run at half load, and a three hour run at 
no load. The results of these tests are of considerable in¬ 
terest, and are here recapitulated. 

The motor which secured first place was a Tony Huber- 
Peugeot single cylinder of no mm. bore by 140 mm. stroke. 
In this motor the charge is precompressed between the 
piston and a sliding plate between the bottom of the cylinder 
and the top of the crank case, with a stuffing box bearing 
for a connecting rod to slide in. The charge is drawn into 
this compression chamber through a suction operated poppet 
valve and is transferred to the combustion chamber through 
a transfer passage and port in the cylinder wall which is 
opened by the piston toward the end of its down stroke. 

The second motor was the Legros, illustrated in Fig. 275. 
This was a two cylinder motor of 100 mm. bore and 120 mm. 
stroke. The test results are given in the following table: 


TWO STROKE MOTOR TESTS. 


Motor. Peugeot. 

Number of cylinders. 1 

Cylinder bore, mm. no 

Piston stroke, mm. 14° 


Legros. 

2 

100 

120 


FULL LOAD TESTS. 


Average speed, r. p. m.1,401.3 967.5 

Average power, h. p. 12.86 12.25 

Total consumption in gallons per hour. 1.91 2.29 

Consumption in pounds per horse power hour. 0.75 1.07 

Weight of motor, pounds. 163.4 360.8 

Weight p. h. p., pounds. 12.7 29.5 











430 


THE TWO STROKE CYCLE MOTOR. 


HALF LOAD TESTS. 

Motor. Peugeot. Legros. 

Average speed, r. p. m.1,427.6 921.25 

Average power, h. p. 6.40 5.40 

Total consumption per hour, gallons. 1.65 1.53 

Consumption per horse power hour, pounds. 1.49 1.6 

NO LOAD TESTS. 

Average speed, r. p. m. 900 920 

Total consumption per hour, gallons. 1.09 1.45 

The piston speeds at maximum output were 1,300 feet per 
minute in the case of the Peugeot, and 930 feet per minute 
in the case of the Legros motor. The corresponding products 
of the mean effective pressures and mechanical efficiencies 
(v p ) were 41.5 and 35 pounds per square inch, respectively. 

A very' complete series of tests of an ordinary two stroke 
motor was made by Dr. W. Watson, of London, and re¬ 
ported by him in a paper read before the Institution of 
Automobile Engineers. The engine was a single cylinder 
one with a bore and stroke of 2> l A inches each. It employed 
crank case compression and was of the three port type. The 
crank case inlet port remained open during 82 degrees of 
crank travel, the transfer port during 97 degrees, and the 
exhaust port during 122 degrees. This engine at 1,506 r. p. m. 
developed 4.16 brake horse power, with a fuel consumption 
of 0.76 pound of gasoline per indicated horse power-hour. 
The mean effective pressure at 1,500 r. p. m. was 53 pounds 
per square inch. At 900 r. p. m. it was 65 pounds per square 
inch. The ratio of air to gasoline at this speed was 12.13 
to 1, and 17 per cent, of the charge supplied to the cylinder 
escaped through the exhaust valve. The loss of charge was 
much greater at low speeds. The compression in this motor 
was between 62 and 65 pounds per square inch at all speeds. 

Air Cooling—Air cooling undoubtedly involves greater diffi¬ 
culties when applied to a two stroke motor than when 
applied to a four stroke motor, owing to the fact that the 
interval between successive explosions is so much shorter in 
the former. Besides this, there is liability of trouble from 
distortion of the cylinders, due to the fact that the exhaust 
side will get hotter than the inlet side, and it is impossible 
to so distribute the cooling effect that the temperature shall 
be fairly uniform all around the cylinder. Notwithstanding 
these difficulties, the problem has been satisfactorily solved, 
as is shown by the fact that one manufacturer, the Chase 
Motor Truck Co., of Syracuse, N. Y., has built three-cylinder, 
air cooled, two stroke motors for automobile work for several 








THE TWO STROKE CYCLE MOTOR. 


431 


years. A cross sectional view of one of these motors is 
shown in Fig. 277* This motor is of the three port type and 
has a bore and stroke of 4% inches each. It will be noticed’ 
that the cylinders are placed a considerable distance apart, 
evidently to minimize the cross radiation effect. Both the 
cylinder and the cylinder head are covered with high cooling: 



Fig. 277.—Chase Air Cooled Two Stroke Motor. 

flanges, and a current of air is drawn over the cylinders by 
means of a flywheel fan. 

Constructional Details—In a two stroke motor it is abso¬ 
lutely necessary to pin the piston rings, for the reason that 
if the rings were free and should turn in their grooves so 
the cut came opposite one of the ports, the ring would bo 
caught by the edge of the port and broken- In. order, to 
































































432 


THE TWO STROKE CYCLE MOTOR. 


prevent trouble from portions of the ring being caught in the 
ports, it is advisable to limit each section of the port to an 
arc of 35 degrees. The presence of the ports makes the 
finishing of a two stroke cylinder a particularly delicate job, 
and grinding is by far the best finishing method. If crank 
case compression is employed some care must be given in 
the design to insuring absolute gas tightness at all points of 
the case. The separate parts of the case are put together 
with paper gaskets between, which are glued to one of the 
parts with shellac. Some manufacturers, however, prefer to 
grind the abutting surfaces on a surface plate, and thus do 
away with the necessity of packing material, which is always 
more or less bothersome when repairs become necessary. 
This method is particularly applicable where the crank case 
is made in halves with a joint in the horizontal plane through 
the crankshaft axis. The bolts for holding the two parts 
together and for securing the cylinders to the crank chamber 
should be placed somewhat closer than is customary in four 
stroke motors. 


CHAPTER XXI. 


MOTOR TESTS. 

There are, essentially, two methods of testing a gasoline 
motor, viz., by means of a registering pressure indicator and 
by means of an absorption dynamometer. We will take up 
these two methods in succession, but will first briefly de¬ 
scribe a method of determining the compression pressure and 
the explosion pressure in the individual cylinders of a multi¬ 
cylinder engine. 

It is sometimes attempted to determine the compression 
pressure in a cylinder by means of a pressure gauge con¬ 
nected to it directly, by turning the engine over by hand. 
This method is very unsatisfactory, however, owing to the 
fact that when the engine is turned over slowly a slight 
leak may allow a great deal of the charge to escape, and 
thus the compression to drop much below its normal value at 
running speeds. Besides, when the cylinder walls are cold, 
a good deal of the heat generated by the compression will be 
absorbed by the walls, and this cooling effect will also reduce 
the compression pressure. 

The Power Indicator—A device designed to enable one to 
determine the correct compression and explosion pressures is 
known as the Gibson power indicator, and is illustrated in 
Fig. 278. This instrument incorporates a pressure gauge indi¬ 
cating up to 300 pounds or more per square inch. At the bot¬ 
tom of the gauge proper there are a number of fittings, in¬ 
cluding an adjustable leak valve A, a check valve B and a 
four-way (for a four-cylinder motor) hand operated cock C. 
The action of the instrument will be readily understood by 
reference to the cut. At each explosion in the cylinder with 
which the gauge is put in communication by the four-way 
cock, a small quantity of burnt gases passes through the 
check valve B into the tube of the gauge, so that in an in¬ 
stant the pressure in the gauge is the same as the maximum 
pressure developed in the cylinder. The gases escape from 


433 



434 


MOTOR TESTS. 


the indicator through the leak valve A at a considerably 
slower rate than they can pass to the indicator through the 
check valve B, so that the reading of the indicator will de¬ 
crease with decreasing pressure in the engine, and the pointer 
will return to zero when communication with the engine is 
shut off. By means of the four-way cock the four cylinders 
of a four cylinder motor can be readily tested and any dif¬ 
ferences in their explosion pressures noted. The compres¬ 



sion pressure in any cylinder can be determined by preventing 
ignition in that cylinder while the engine is running. 

In the regular use of the device, when the engine is run¬ 
ning, if the ignition be imperfect and misfires occur the action 
of the gauge hand will be erratic, the hand advancing a con¬ 
siderable distance and then slowly returning to the compres¬ 
sion pressure. If ignition is perfect the beat of the hand will 
be regular and can be adjusted by means of the leak valve. 


























































































MOTOR TESTS. 


435 


Piston Type Indicators—In studying the internal action in 
the cylinders of steam and slow speed gas engines use is 
made of a so-called engine indicator consisting of a small 
pressure cylinder which is placed in communication with the 
working cylinder, and a piston in this small cylinder with a 
piston rod connected through a suitable linkage with an arm 
carrying a pencil at its free end. As the pressure in the 
working cylinder rises and falls the pencil moves up and 
down the side of a small drum which carries a card of white 
paper. The drum is simultaneously rocked back and forth 
around the axis by connection with the crankshaft or some 
other moving part of the engine. The simultaneous motion 
of the drum and of the pencil produces on the card a co¬ 
ordinate diagram with abscissas representing piston travel 
and ordinates representing gas pressure. These diagrams were 
discussed at some length in Chapter III, and the principle of 
the indicator was shown diagrammatically in Fig. 5. Unfor¬ 
tunately these instruments, in spite of the greatest refine¬ 
ments, are practically useless above speeds of 500 revolutions 
per minute. This is due to the fact that the inertia of the 
piston and other reciprocating parts then becomes so great 
as to seriously affect the accuracy of the diagram. 

The Manograph.—The manograph is an optical indicator in 
which the place of the pencil carrying arm is taken by a ray 
of light. This principle was first suggested by Prof. John 
Perry, but apparently the first manograph for the study of 
gasoline motors was made for Prof. E. Hospitalier, of the Ecole 
de Physique et de Chimie Industrielles, of Paris, by J. Carpen- 
tier, the well known instrument maker. It was first used in tests 
of motors exhibited at the alcohol exhibition in Paris in Oc¬ 
tober, 1901. Since that time it has been further developed, 
and at present either it or the Schulze, a similar instrument, 
is used in nearly all automobile engineering laboratories of any 
pretension. 

The construction of the manograph is clearly shown by the 
sectional views Figs. 279 and 280. In these figures A is the 
source of light. In the original instruments an acetylene 
light was used, but this proved rather unsatisfactory, and 
Mr. Gibson, the importer of the Hospitalier instrument 
in this country, substitutes therefor a hand-operated electric 
arc, a suitable resistance being connected in circuit with the 
arc so as to insure steady operation. A pinhole on the fo¬ 
cusing slide B allows a practically parallel ray of light to pass 
into the interior of the instrument and fall upon the total re- 


136 


MOTOR TESTS. 


flection prism C. This prism reflects the ray of light onto the 
small concave mirror D, which reflects it onto a polished 
glass screen E, which forms one head of the box of the in¬ 
strument. The mirror D is located in the centre of the oppo¬ 
site head of the box, and the length of the box is such that 
the focal point of the mirror coincides with the mat surface 



Fig. 279. —Part Sectional View of Hospitalier Manograph. 


of the ground glass screen. The resultant point of light is so 
small that its intensity is little, if any, less than the intensity 
of the light at the pinhole, and the motions of the point of 
light on the plate are easily followed by the eye, even in 
broad daylight. 

Construction of the Instrument —The mirror is given a rock¬ 
ing motion in a vertical plane proportional to the fluctuations of 
the pressure within the engine cylinder and a rocking motion in a 


































































MOTOR TESTS. 


437 


horizontal plane proportional to the travel of the piston. Refer¬ 
ring to Fig. 280, back of the mirror is a metal plate 1-2-3 in 
the form of a right angled triangle. This plate rests upon 
three supports, at its three corners, respectively. Point 1, at 
the right angle, is stationary in space. Point 2 is displaced 
perpendicularly to the plane of the mirror in proportion to 
the pressure, and point 3 is displaced proportionally to the 
distance traversed by the piston. Point 2 of the plate rests 
against the end of a steel rod F, the other end of which bears 

against the centre of a steel diaphragm G clamped between the 

cap H and plug I. From the plug I connection is made to the 

engine cylinder through a small bore copper tube, whereby the 

pressure acting on the cylinder is communicated to the dia¬ 
phragm chamber G. The diaphragm is of such construction that 
the deflection of its centre is as nearly as possible proportional 
to the pressure acting upon it, and sometimes consists of two 
discs of different grades of steel whose errors compensate each 
other. Thus the point 2 of the plate back of the mirror will 
be moved a distance proportional to the pressure in the cylinder. 

Point 3 of the plate back of the mirror rests against the steel 
rod J, the opposite end of which rests against one arm of a 
double armed lever K, whose other arm connects through 
a short connecting rod L to a small crank M. The shaft 
of crank M carries a spur pinion meshing with another spur 
pinion N, both pinions having the same number of teeth. Pinion 
N is connected through a flexible shaft O with the crank¬ 
shaft of the engine under test. By means of the crank and 
connecting rod mechanism LM, the linear motion of the pis¬ 
ton is reproduced on a reduced scale and impressed upon the 
sliding rod /. The mirror is constantly pressed against its 

three supports by the spring The reciprocating motion of 
the rod / must be absolutely in phase with the motion of the 

piston, and in order that this relation may be attained an 

adjusting device is provided which is arranged as follows: 
The shaft of pinion N has a bearing in the disc P concentric 
with the pinion M. This disc P is provided with worm teeth 
with which meshes a worm provided with a thumb-wheel Q. 
By operating the thumb-wheel Q the disc P can be moved 
around its centre whereby the angular relation of crank M 

to the engine crankshaft is changed. When the motion of 
the pin J is brought into absolute phase with that of the pis¬ 
ton, the point 3 of the plate back of the mirror will be moved 
in direct proportion to the motion of the engine piston. The 
beam of light reflected by mirror D will then travel up and 


438 


MOTOR TESTS. 


down on the screen E proportional to the pressure in the en¬ 
gine cylinder and transversely across the screen in propor¬ 
tion to the motion of the engine piston. Owing to the per¬ 



sistence of vision the diagram appears as a continuous line 
which may be examined conveniently or traced with a pencil 
on tracing paper. By putting a sensitized plate or film in the 
place of the ground glass plate, the diagram may be photo- 




































































































































MOTOR TESTS. 


439 



Fig. 2S1. Single Cylinder JTospitalier Manograph. 










440 


MOTOR TESTS. 


graphed, and the photograph then used the same as the ordi¬ 
nary indicator diagram. 

Sources of Error —There are several sources of error in 
this instrument which it may be well to point out here. In 
the first place, the piston motion is reproduced accurately 
only if the ratio between the length of connecting rod L and 
the throw of crank M is the same as the ratio of the engine 
connecting rod length to engine stroke. The length of con¬ 
necting rod L in the instrument should therefore be changed 
in accordance with the ratio of connecting rod length to 
length of stroke in the particular engine under test. This, how¬ 
ever, is impractical in actual work with the instrument. Another 
source of error resides in the flexible shaft and its connec¬ 
tions, which will twist slightly in operation, and this twist will 
vary with the speed of the motor. This error can be cor¬ 
rected to a certain extent by means of the thumb-wheel ad¬ 
justment Q. A further source of error lies in the copper tube 
connecting the engine cylinder with the diaphragm chamber. If 
this tube is made very small in diameter and of considerable 
length, its resistance to the flow of the gases may cause a dif¬ 
ference in pressure to exist between the cylinder and the 
diaphragm chamber. On the other hand, if the tube is made 
of fairly large bore it has an appreciable effect on the capacity 
of the compression chamber, and if it is made very short 
a great deal of heat will be transmitted to the diaphragm 
chamber, which may affect the accuracy of the diaphragm. 
The Hospitalier manograph is furnished with a copper tube 
of 14 inch bore and about 3 feet in length. When a very 
short length of tube is used it is necessary to water-cool it. 
A single cylinder Hospitalier manograph complete is shown 
in Fig. 281. 

Phase Adjustment —The first operation in making a mano¬ 
graph test, after the instrument has been set up and con¬ 
nected to the engine, consists in adjusting for phase equality. 
This, as already explained, is done by means of the thumb¬ 
wheel Q. Suppose that the engine is turned over by hand. 
Then during the compression stroke the pressure in the cyl¬ 
inder will rise to near the normal compression, and during 
the following down stroke it will drop practically along the 
same line. If there were no leakage and no interchange of 
heat between the gases and the cylinder walls the compres¬ 
sion and expansion curves would absolutely coincide. As 
a matter of fact, supposing the cylinders to be tolerably 
airtight and to be cold, the compression and expansion. 


MOTOR TESTS. 


441 


curves when the instrument is properly synchronized will 
be substantially as in Fig. 282. That is, during the first 
part of the expansion the curve will drop a little below the 
compression curve, owing to the fact that heat is then ab¬ 
sorbed by the cylinder walls from the gases, while during the 
latter part of the expansion the expansion curve will coincide 
with the compression curve. Therefore, when a diagram 
like Fig. 282 is obtained, the operator knows that the instru¬ 
ment is in phase with the engine. If the expansion curve falls 
much below the compression curve during the first part of 
the expansion is indicates cylinder leakage. When the en¬ 
gine is turned over while the communication between the 
cylinder and the diaphragm chamber is shut off, the spot of 
light describes a straight horizontal line on the screen, which 



Fig. 282.—Compression Diagram Used for Adjusting the 
Phase of the Manograph. 


represents atmospheric pressure, and is known as the atmos¬ 
pheric line. 

Studying Ignition and Combustion —In some investiga¬ 
tions, however, it is advantageous to have the instrument out 
of phase with the engine, as, for instance, when trying to de¬ 
termine the relative rates of combustion of different fuel 
mixtures. This factor is indicated by the inclination of the 
explosion line in the diagram. When the instrument is ad¬ 
justed in phase with the motor the explosion takes place 
while the rod / is stationary or nearly so, and any slight 
variation in the rapidity of combustion, therefore, does not 






442 


MOTOR TESTS. 


appreciably affect the diagram. On the other hand, if the in 
strument is adjusted so as to be 90 degrees out of phase with 
the engine, the piston, and, consequently, the rod J control¬ 
ling the horizontal motion of the beam of light, will move 
at its maximum speed while the explosion occurs, and any 
variation in the rate of explosion is shown on a much larger 
scale. A diagram taken with the instrument out of phase by 
about 90 degrees is shown in Fig. 283. 

Four Cylinder Manographs —For studying a multi-cylinder 
engine one manufacturer recommends the use of a mul¬ 
tiple manograph for taking records from all of the cylinders 
simultaneously. If all of the cylinders are in the proper working 
condition, then the different diagrams taken at the same time should 



Fig. 283.—Diagram Taken with Crank Motion 90 Degrees 

Out of Phase. 

be absolutely identical. If any variations are shown in the diagrams 
it is then a comparatively easy matter to trace the cause of 
the irregularity. On the other hand, if all of the cylinders 
were indicated with a single instrument connected to one 
cylinder after another, the working conditions might change 
during the interval between successive records, so that the 
different diagrams would no longer be directly comparable. 
Fig. 284 shows a four cylinder Schulze manograph complete. 

There is, however, something to be said in favor of the use 
of a single instrument on engines with any number of cylinders. 
No two diaphragms are absolutely identical in their actions and 
when the cards from the different cylinders are taken with 
different instruments, the differences in the diaphragms may be 
responsible for some of the differences in the diagrams. More- 




MOTOR TESTS. 


443 



Fig. 284. —Four Cylinder Schulze Manograph. 

A. source of light (Nernst lamp); B, transmission of crank motion. 


over, by means of bayonet joints the single instrument can be 
very quickly disconnected from one cylinder and connected to the 
next. 

Use of Manograph —One use of the manograph consists in 
determining the indicated horse power. By subtracting the brake 
horse power from the indicated horse power, the internal fricticn 
losses and the mechanical efficiency of the motor are found. 

The area enclosed by the compression and expansion curves 
represents work done upon the piston by the expanding 
gases. In case portions of the exhaust and compression 
curves and the suction curve form a loop, as they usually do, 



Fig. 285. —High Exhaust Back Pressure 

(Shaded area represents negative work.) 





























444 


MOTOR TESTS. 


the area enclosed by them (the shaded area in Fig. 285) rep¬ 
resents negative work and must be subtracted from the posi¬ 
tive work represented by the area enclosed by part of the ex¬ 
haust and compression curves and the expansion curve. In 
order to find the indicated horse power it is necessary to de¬ 
termine the mean height of the diagram. After the mean 
height of the positive work area and the mean height of the 
negative work area have been determined and their difference 
has been taken, the mean effective pressure is found by mul¬ 
tiplying this difference.by the scale of the diaphragm, which 
is generally expressed in pounds per square inch per inch. 
The power developed by the cylinder from which the dia¬ 
gram was taken may then be determined from the formula 
_ M. P. P. X W; 2 X 2 IX n _ tt b 2 / X AT. E. P. X n . 

^ ' ' 4X4X12X33 000 — 3 r68 000 



Fig. 286.—Diagram Taken from De Dion y / 2 Horse¬ 
power Motor. 

where M. E. P. is the mean effective pressure as determined 
from the diagram; b the bore of the cylinder in inches; l the 
length of stroke in inches and n the number of revolutions per 
minute. The speed of the engine must be taken at the same 
time as the diagram, if the indicated horse power is to be de¬ 
termined. 

The other and, perhaps, the most important use of the 
manograph consists in determining faults in the operation of 
the motor due to defective valve setting, ignition, carbure- 
tion, etc. If the engine works under the proper conditions it 
gives a diagram very closely approximating the ideal diagram. 











MOTOR TESTS. 


445 


as discussed in Chapter III. From the known data of the 
cylinder under test such a diagram may be drawn, and the 
diagram obtained by means of the instrument may then be 
compared with it. 

A very good diagram is shown in Fig. 286. This was taken 
by Joseph Tracy with a Schulze manograph, using a Nernst 
lamp as the source of light, from a single cylinder 3T2 h. p. 
De Dion motor. The maximum pressure indicated is 218 
pounds per square inch, and the mean effective pressure 59.8 
pounds per square inch. The diagram was taken at 1,322 
r. p. m. of the motor, and the indicated horse power is equal 
to 2.4. 

Fig. 287 shows a diagram taken with the Schulze manograph 



Fig. 287.— Diagram Taken from Franklin Air Cooled 

Motor. 

from a Franklin air cooled, four cylinder motor with auxil¬ 
iary exhaust ports. This also corresponds closely to the ideal 
form. A striking feature of this diagram is the sharp devia¬ 
tion of the expansion line at the point where the auxiliary 
exhaust port begins to open. However, Fig. 286 shows that 
a quite similar effect can be obtained with the ordinary me¬ 
chanically operated exhaust valve, provided it is opened early 
enough. An example of the effects of an obstructed exhaust 
and consequent high back pressure during the exhaust stroke 
is seen in Fig. 285. This diagram is taken from a paper read 
before the Society of Automobile Engineers, descriptive of 
some tests made at Cornell University, and it is believed that 
the high back pressure was due to the fact that in the tests 









446 


MOTOR TESTS. 



Fig. 288.—Diagram Showing Pre-ignition. 


a muffler was connected to the engine very close to the ex¬ 
haust valves. 

A diagram indicating too early ignition (pre-ignition) is 
shown in Fig. 288, and a diagram indicating late ignition in 
Fig. 289. 

Fig. 290 shows a diagram taken from one cylinder of a six 
cylinder motor in which, owing to the firing order and the 
arrangement of the exhaust manifold, there was a transference 
of dead gases from one cylinder to another. As already ex¬ 
plained in a previous chapter, the little loop at the left hand 
side indicates this blowing over. The explosion and expansion 
lines indicate a slow burning mixture, which may be due either 



Fig. 289.— Diagram Showing Too Late Ignition. 








MOTOR TESTS. 


447 


to excessive dilution of the charge with dead gases, to poor 
carburetor adjustment, or to both. 

Manographs are usually furnished with two different sets 
of diaphragms, one a high pressure set for indicating the cyl¬ 
inder while in regular operation, and the other a low pressure 
set for studying the action of the valves, with the ignition 
cut off. The Schulze manograph is furnished with a dia¬ 
phragm giving a scale of 3 mm. per kilogram per square cen¬ 
timetre (one inch per 120 pounds per square inch) for the 
regular working stroke, and a diaphragm giving a scale of 
10 mm. per kilogram per square centimetre (one inch per 36- 
pounds per square inch) for compression diagrams. 

The ground glass .screen of the manograph is generally so 
arranged that a photographic plate holder can be inserted in 
its place. However, Mr. Gibson has devised a more convenient 
method for photographing the diagram. This consists in at¬ 
taching a Premo film pack to the frame of the screen. Adjust¬ 
ments on the engine are made while the ground glass screen 
covers one end of the instrument so that the diagram can be 
seen upon it, and when the proper shape of diagram has been 
obtained the frame is shifted to one side so that the film pad 
covers the end of the box. The taking of a photographic dia¬ 
gram is then the work of an instant. A photographic shutter 
inserted between the source of light and the mirror insures 
that the pressure variations of only one cycle are inscribed' 
upon the photographic plate, so that the diagram gives an ac¬ 
curate record of that cycle. In conjunction with this shutter 
a very strong source of light is required, since the spot of light 
travels over the sensitized plate only once and considering the- 
spot of light to be 1-64 inch in diameter it can be easily 
figured out that the light has less than one ten-thousandth of 
a second to act. 



Fig. 290.—Diagram Showing Slow Combustion of Diluted* 

Charge. 







448 


MOTOR TESTS. 


Dynamometers —The manograph is of use chiefly in deter¬ 
mining improper gas flow, which as a rule is due to defects in 
construction. It is mainly a laboratory instrument, and is used 
only on engines while being developed experimentally. In con¬ 
nection with such problems as the best valve timing, sizes of 
valves and sizes and location of ports in two cycle motors, the 
manograph may render valuable service. However, the chief 
aim in motor development is always increased brake horse power, 
and for this reason the testing device most generally used and 
which is applied to every motor turned out, is some form of 
power brake or dynamometer which enables the tester to deter¬ 
mine the power output directly. 

An apparatus for determining the power output of a motor is 
known as a dynamometer. Dynamometers are divided into two 
classes, viz., absorption dynamometers, which absorb and dissi¬ 
pate the energy output of the motor, and transmission dynamom¬ 
eters, which are inserted between the motor and its load and do 
not absorb any part of the motor output. All of the dynamom¬ 
eters which are practically employed for testing automobile 
motors are of the first kind, i. e., absorption dynamometers. The 
most primitive form is known as the Prony brake. 

The Prony Brake—This consists of a pair of blocks which 
can be applied to the rim of the flywheel or of a pulley secured 
to the flywheel or crankshaft, with adjustable pressure. The 
friction between the blocks and the pulley will absorb the output 
of the motor and convert it into heat. The blocks are secured 
to an arm whose other end rests upon the platform of a beam 
scale or is attached to a spring scale. Referring to Fig. 291, 
when the motor rotates in the direction indicated by the arrow, it 
will cause the scale to indicate a pressure. This pressure is due 
partly to the friction of the blocks on the drum and partly to the 
weight of the arm. Since it is desired to determine the friction 
alone, either a correction must be made for the weight of the 
arm resting on the scale or else the effect must be eliminated by 
providing a counterweight on the side opposite the arm, so the 
brake balances around the centre of the bore of the brake blocks. 

Let the radius of the drum be r and the distance of a vertical 
line through the point of contact on the scale, from the centre 
of rotation, D. Then, when the motor is running and the scale 
indicates a pressure W, the friction at the circumference of the 

drum is evidently — W. If the speed of revolution of the motor 

r 

be n, then the rubbing speed at the rim of the pulley or flywheel 
is 2 7T rn inches per minute, and the power dissipated in fric¬ 
tion is 

• 


449 


MOTOR TESTS. 



Fig. 291.—Prony Brake. 


D 


2 7 T r 71 


H. P. =- W X 

r 12 X 33,000 


D IV n 
63,025 


The Prony brake is not a very practical instrument. Some dif¬ 
ficulty is experienced in carrying off the heat generated quickly 
enough. This problem is generally solved by using a pulley with 
a rim with inwardly turned flanges on both sides and directing 
a stream of water into the pulley. The friction at the circum¬ 
ference changes almost continually, owing to wear and to the 
expansion due to heating, and for this reason it is not possible 


to extend Prony brake tests over 
conditions. 

The Rope Brake—An im¬ 
proved form of friction brake 
is the rope brake. The type 
illustrated in Fig. 292 is recom¬ 
mended by the American So¬ 
ciety of Mechanical Engineers. 
In this brake the friction tends 
to lift the weight, and is 
measured by the difference be¬ 
tween the weight Wi and the 
scale reading when the pulley 
is turning. The load is auto¬ 
matically adjusting, which 
makes this brake better suited 
to accurate tests than the 
Prony brake. The calculation 
of the power obtained in a 
test with this brake is similar 
to the calculation for the 
Prony brake. The difference 
between the weight W and the 


considerable periods under fixed 








































450 


MOTOR TESTS. 


indication of the scale is the equivalent friction at the centre 
of the rope, and the power may therefore be found in the most 
convenient way by supposing the friction surface to be at the 
centre of the rope. For use where the distance of the pulley 
or flywheel from the floor is not sufficient to allow' of the use 
of a brake of the type illustrated, the Society of Mechan¬ 
ical Engineers recommends a similar type comprising a double 
armed lever hung from the ceiling. The weight is hung from 
one arm of the lever and to the other arm is attached the rope, 
which, after being wound around the pulley or flywheel, is at¬ 
tached to a spring scale anchored to the floor at some distance 
from the pulley. 

The Fan Dynamometer—The fan dynamometer, which was 
first introduced in 1902 by Colonel Renard, of the firm of Pan- 
hard & Levassor, of France, consists merely of a wooden beam 
bolted to the end of a horizontal shaft and carrying two square 
discs of aluminum, which may be moved farther in toward the 
shaft or farther out from it, as desired. Small sizes of the 
dynamometer may be mounted directly upon the crankshaft of 
the engine to be tested, but in the larger sizes it is preferable 
to fix the beam to a special shaft which is mounted in ball bear¬ 
ings in a bearing housing bolted to a supporting frame. At its 
free end this shaft is provided with a member of a universal 
joint, so it can be connected to the crankshaft by means of an 
intermediate shaft and two universal joints. The advantages of 
a dynamometer of this type, as compared with a Prony brake, 
are obvious. The resistance does not change as the test pro¬ 
ceeds, and the problem of keeping the device from overheating 
does not arise. 

As the dynamometer was originally constructed by Colonel 
Renard the beam was made of ash and provided with two alu¬ 
minum discs secured by two bolts each. To allow of the distance 
of the discs from the centre line of rotation being varied, each 
arm of the beam was drilled with eleven bolt holes, equidistant 
from each other. These eleven holes admitted of ten different 
positions of the discs. Ten intermediate positions were obtain¬ 
able by means of extra holes in the discs, as shown in Fig. 293. 
The distance between holes was 5.5 centimetres (2.2 inches), the 
total length of the instrument 52.8 inches, and the complete 
weight 26.4 pounds. 

The resisting torque of a fan dynamometer is proportional to 
the square of the angular velocity, and the power absorbed to 
the cube of the angular velocity. Both the resisting torque and 
the power absorbed are also proportional to the density of the 


MOTOR TESTS. 


451 


atmosphere, but the variations of this factor are negligible in 
practical work. This type of dynamometer is particularly valu¬ 
able for adjusting a motor. A position for the disc is first found 
experimentally which allows the motor to run at about its normal 
speed. Then the tester begins to make his adjustments, and any 
improvement in the adjustments is indicated by an increase in 



Fig. 293.—Renard Fan Dynamometer. 

the speed of the motor. When it is desired to rapidly test a con¬ 
siderable number of motors of the same size, the best position 
of the fan discs is first determined. All of the motors should 
run at the same angular speed, and if any motor fails to come 
up to the speed shown by the others, it is a sign that it is de¬ 
fective in some way. 

It has been proved both by calculation and by experiments that 

















































452 


MOTOR TESTS. 


for fans of proportional dimensions throughout the power co¬ 
efficient varies as the fifth power of the ratio of linear dimensions. 
This law was verified by means of experiments with two instru¬ 
ments the linear dimensions of which bore to each other the 
ratio of 2.2. The fifth power of 2.2 being equal to 51.54, the co¬ 
efficient of the larger dynamometer should be equal to that of 
the smaller one multiplied by 51.54. In spite of the inevitable 
experimental errors and inaccuracies of construction, this rela¬ 
tion was found to hold very well for all positions of the discs 
on the beam. 

When a dynamometer is constructed with dimensions of re¬ 
volving parts bearing to each other relations different from 
those given in the table below, it is necessary to calibrate the 
instrument by means of some standard form of dynamometer. 
Then, when it is desired to build larger or smaller instruments 
and adhere to the same proportions of parts, all that is neces¬ 
sary is to multiply the power factor found for the first instrument 
by the fifth power of the ratio of the linear dimensions in the 
original and new instruments respectively. Colonel Renard found 
a difference of less than one-half of 1 per cent, in the power 
indications of a number of dynamometers of the same size con¬ 
structed with ordinary care, and he expressed a belief that with 
all-metal construction the difference would be smaller. The 
relative dimensions of the parts of the Renard fan dynamometer, 
expressed as a function of the modulus a (distance between ad¬ 
jacent bolt holes), are as follows: 


Distance between holes. a 

Length of beam. 24 a 

Thickness of beam. a 

Width of beam. 2 a 

Side of square discs. 60/11 a 

Thickness of discs. 4/55 a 

Diameter of bolts. 10/550 

Diameter of shaft. 40/550 

Number of holes in each arm of beam. 11 

Distance from the centre to the outermost hole. 11 o 


The maximum power for which a dynamometer of this kind 
may be used is determined by the limit of speed at which the law 
of proportionality of the air resistance still holds and by the 
limit of fibre stress in the beam. In experiments a peripheral 
velocity of 328 feet per second has been obtained, but it is possi¬ 
ble that the law of the square of the speed holds at still higher 
velocities. The fibre tension in the beam, on the other hand, is 
limited to 1,400 pounds per square inch, which with selected ash 












MOTOR TESTS. 


453 


wood gives a safety factor of 7 to 8. By combining these two 
conditions the limits of the range of a given dynamometer are 
easily determined, as regards both angular speed and power. It 
can easily be demonstrated that all fans of proportional dimen¬ 
sions have the same factor of safety when running at the same 
peripheral speed. This equality of the safety factor applies not 
only to the beam but also to the discs, the bolts and the shaft. 



Fig. 294. —Tracy Fan Dynamometer. 

An improved type of fan dynamometer is made by Joseph 
Tracy, of New York, and is illustrated in Fig. 294. It has a 
metal base and is provided with a tachometer and power indi¬ 
cator, the instrument having two scales, reading in revolutions 
per minute and horse powers respectively. In the regular use of 
a fan dynamometer the fan must be covered with a heavy wire 
screen for the sake of safety. 









454 


MOTOR TESTS. 


If a steel beam is used in a fan brake, of a section compara¬ 
tively narrow and long, the power absorbed will be substantially 
proportional to the area of the movable plates and to their 
distance from the axis of rotation. Furthermore, if the beam 
is slightly offset at the centre where it is secured to the shaft, 
so that the two plates are both in the same plane through the 
axis of the shaft, then when the beam is drilled with equidistant 
holes the increase or decrease in the centre distance of the plates 



voiutio n «y per' Minu fe. 


Fig. 295. —Fan Dynamometer Calibration Curve. 


from the axis of rotation as the plates are moved farther out or 
in will be proportional. The theory of such an instrument, there¬ 
fore, will be somewhat simpler than that of the wood beam fan. 
However, the relation between the power absorbed and the 
position of the plate on the beam is a very complicated one, 
and it is usual practice to determine the power absorbed by such 
a dynamometer for the different positions of the plates experi¬ 
mentally and prepare a set of curves from which the horse power 
absorbed at any given speed of revolution and any given position 
of the plates can be read off directly. Experiments conducted 

























MOTOR TESTS. 


455 


by Prof. W. Morgan and E. B. Wood, of England, have shown 
that the horse power absorbed varies directly as the third power 
of the number of revolutions per minute, and the calibration 
curves shown in Fig. 295 bear out this result. The latter apply to 
a fan dynamometer of the type described above, with steel beam, 
with plates 12 inches wide by 8 inches radial depth. The figures 
on the individual curves denote the radial distances from the 
axis of rotation to the geometrical centre of the plates. 



Electric Cradle Dynamometer—Perhaps the most practical 
device for making accurate horse power tests on a motor is 
the electric cradle dynamometer. This consists of a dynamo 
of the circular field frame type. The field frame, instead of 
being rigidly bolted to the base or cast integral therewith, is 
supported in two large radial ball bearings on pedestals rising 
from a cast iron base. The armature shaft is connected to the shaft 
of the motor to be tested through a flexible coupling. A sketch 
of such a dynamometer is shown in Fig. 296. When the 
armature is rotated by the gasoline motor it exerts a turning 
effort on the field frame A and tends to carry the latter around 
with it. Two radial arms extend from the field frame, one, B, 
for measuring the pull exerted by means of a scale, and the 
other, C, for balancing the former. Current is induced in the 
armature of the generator, the same as in the ordinary dynamo, 
and this may be used in lighting up a bank of incandescent 
lamps or it may be dissipated in a water rheostat. The pressure 
exerted by the lever arm on the scale is equivalent not only to 















456 


MOTOR TESTS 



Fig. 297. —Electric Cradle Dynamometer in A. C. A. Laboratory. 














MOTOR TESTS. 457 

the electromagnetic reaction between the armature and field 
frame, but also to the armature bearing friction and the commu¬ 
tator brush friction, and the only item not included in this 
pressure is the friction in the ball bearings supporting the field 
frame, which is negligible, however. The method used in con¬ 
nection with such a dynamometer for calculating the horse power 
is exactly the same as that used in connection with the friction 
brake, and the same formula can be used, i. e., 

D IV n 
H. P. = -z - 

63,025 

As in the cases of all other dynamometers, torque and speed 
measurements must be made simultaneously, and for this reason 
a tachometer is generally mounted on top of the dynamometer. 

Fig. 297 shows a dynamometer installation at the laboratory of 
the Automobile Club of America by the Diehl Manufac¬ 
turing Co., of Elizabeth, N. J. The installation comprises a 
large cast iron base with adjustable standards to take motors of 
all sizes, an electric cradle dynamometer and a switchboard. The 
dynamo has a capacity of 64 horse power at 2,000 revolutions, 
but will stand an overload of 90 horse power at 1,500 revolutions 
per minute for short periods of time. The dynamometer arm in 
this case is 3 feet long, measured from the centre of the dynamo, 
and is provided with a knife edge, to insure absolute correctness 
of the length of the effective lever arm. A feature of the dynamo 
is that it is provided with commutating poles, the object of which 
is to prevent sparking at the commutator under all conditions of load 
without adjustment of the brushes. The dynamo is also capable 
of operating as a motor, and by means of a starting rheostat on 
the switchboard can be used to start the engine. 

The electrical energy developed during a run is dissipated in 
the resistance boxes, and in case of necessity a water rheostat 
may be used in addition. These resistance boxes are controlled 
by individual switches mounted on the lower part of the switch¬ 
board, and the load can be varied from 5 amperes at 220 volts 
(slightly over one horse power) up to the maximum capacity of 
the machine. In the centre of the switchboard, Fig. 298, are 
rheostats for coarse and fine adjustments, respectively, the two 
being connected in series. It is claimed that the resistance can 
be so finely adjusted that it is possible to keep the beam of the 
scale in a floating position. A voltmeter and an ammeter fitted 
to the sides of the switchboard on top give indications from 
which the output of the dynamo can be calculated, and if the 
efficiency of the dynamo at different outputs is known, the input. 



158 


MOTOR TESTS. 



Fig. 298.— Switchboard of Dynamometer Installation. 

which is equal to the output of the motor under test, can also 
be ascertained electrically, thus furnishing a check on the 
mechanical determination of the power. 

Hydraulic Dynamometers—Another type of dynamometer 
used in testing automobile motors is the hydraulic dynamometer. 
One of these instruments, as used in the plant of the American 
Locomotive Works, Providence, R. I., is shown in Fig. 298. The 
instrument comprises a stator and a rotor. The rotor consists of 
four steel discs, 40 inches in diameter and R* inch thick, securely 
keyed to a substantial shaft driven from the motor under test 
through a universal jointed shaft and multiple disc clutch. The 
brake discs revolve in four separate compartments within the 
casing, but clear the walls in all directions by inch, the only 
metallic contact between the rotor and the stator being in the 













MOTOR TESTS. 


459 



Fig. 299.—Alco Hydraulic Dynamometer. 
















































































































460 


MOTOR TESTS. 


bearings which carry the shaft. The casing is free to revolve 
through a small angle and its motion is practically frictionless, 
owing to the special method of mounting it. An arm projects 
from the periphery of the casing, the end of which is linked to a 
steel beam supported on knife edges by a pillar in the same man¬ 
ner as the beams of platform scales, the whole forming a com¬ 
pound lever which is balanced by means of an adjustable weight. 
The scale beam is graduated to read pounds, and by means of a 
vernier, tenths of pounds, a sliding weight being moved along 
the beam by means of a screw and hand wheel so arranged as 
not to disturb the balance. The leverage of the beam is so cal¬ 
culated that at 1,000 revolutions per minute one pound represents 
one horse power. Thus the horse power can be read off directly at 
1,000 revolutions per minute, and at other speeds it can be obtained 
by simply multiplying the weight by the speed and dividing by 
1 , 000 . 

When it is desired to make a brake reading, the rotor is first 
engaged by means of the multiple disc clutch. Water is then 
admitted to one of the erjd compartments near the shaft, and its 
flow regulated by an ordinary globe valve. Centrifugal force 
carries the water out to the periphery of the disc, and creates a 
high pressure, the water then serving as a friction medium be¬ 
tween the steel disc and the walls of the compartment. As a 
result, the casuig tends to rotate with the disc in direct proportion 
to the friction, and hence to the torque. Whatever friction results 
from the rotation of the shaft in the bearings of the brake casing 
has a tendency to revolve it in the same direction; therefore, the 
entire power developed by the motor is accounted for by the 
weighing apparatus, which counteracts the tendency of the casing 
to rotate. 

If one disc or compartment is not capable of absorbing the 
total power of the motor, the compartment at the other end of 
the casing can be brought into service, and the water in these 
compartments may be transferred to the inner compartments by 
means of by-pass valves, thus bringing all four discs into use 
in absorbing the maximum power of which the brake is capable. 
By varying the quantity of water in the compartments a very 
flexible control is obtained. This is done by regulation of the 
inlet and outlet, or drain, valves. The mechanical energy of the 
motor is converted into heat inside the brake and is carried 
off by the water, the temperature of which is usually kept below 
150 degrees Fahr. The brake water, as well as the jacket 
water, can be cooled in a radiator and used over and over. 


MOTOR TESTS. 


461 


Fuel Consumption Tests —In developing a motor experi¬ 
mentally as well as in determining the effect of different carbu¬ 
retors, carburetor adjustments or ignition equipments, it is often 
desirable to make a test of the fuel consumption of the motor. 
This is readily accomplished by means of a graduated fuel tank, 
as illustrated in Fig. 300. Tank A is made of sheet metal, and 
is of small but uniform cross section, so that if a certain amount 
of fuel is drawn it will make quite a difference in the level of 
the fuel. The supply pipe leading 
to the carburetor leaves the tank 
at the bottom, and an extra outlet 
with a nozzle B may be provided, 
which is convenient for filling a 
priming can. The glass tube C is 
connected to the tank at its upper 
and lower ends, and is graduated 
in pounds by means of a paper or 
linen scale pasted to it, or a wooden 
scale firmly fastened back of it. 

The difference between the scale 
readings at the beginning and end 
of a run gives the fuel consump¬ 
tion during the run. Under ordi¬ 
nary conditions the lever cock D 
inserted in the feed pipe of the 
graduated tank remains open, the 
fuel then merely passing through 
the tank, but when a fuel test is 
to be made this cock is closed and 
at the same time the tank is vented 
so as to prevent the formation of 
a vacuum as the level of the fuel 
in the tank drops. 

A distinction must be made be¬ 
tween testing motors in the ex¬ 
perimental development of the de¬ 
sign and testing the regular 

factory product. In the former case great accuracy is essential, 
and methods may be employed which would be impractical in the 
regular testing of the stock product. 



Pig. 


300.—Graduated 
Fuel Tank. 


“Running In” —After the motor is assembled it is first sub¬ 
mitted to a process known as “running in,” which consists in 
limbering up the bearings by running it from a separate source 
of power for from two three hours. With the small clearances 


































462 


MOTOR TESTS. 


allowed for in the bearings, pistons, etc., the motor, after as¬ 
sembling, is usually very stiff, and may require as much as 50 to 
60 per cent, of its normal power output to run it. All of this 
power is converted into heat, and in order that the heat may 
be carried off rapidly from the bearings where it is generated, 
and that cutting at the bearing surfaces may be prevented, the 
motor is flooded with oil, the crank case being filled with a 
heavy grade of oil to a much greater height than in regular 
operation. For this test the motor is not piped up and the 
caps over the valves are left out. Some of the oil will be forced 
out through the exhaust plug openings and spark plug holes, etc., 
and in order to prevent waste of expensive oil it is a good plan 
to cover the motor with a sheet metal hood and place an inclined 
pan under it, so that all oil blown out will be caught and col¬ 
lected in a tank placed underneath the lower end of the inclined 
pan. 

After the bearings have been run in the motor is connected 
up with a gasoline tank, ignition outfit, water supply, etc. Gaso¬ 
line tanks and ignition outfits are usually secured to the wall or 
to standards near each testing stand. If water is plentiful it 
may be taken from the regular factory supply and discharged into 
the sewer after passing through the cylinder jackets, but if it 
is scarce a large cooler may be set up with a blower, for cooling 
the water and thus using it over and over. 

After the connections have been made a little gasoline is 
injected into each cylinder through the pet cocks, the pet cocks 
are closed, the spark lever is set for a late spark, the throttle is 
opened about half way, and the engine is started by means of 
the crank. Most likely it will start off on one of the first 
attempts. The tester then immediately advances the spark and 
cuts down the throttle opening, if necessary, to obtain a speed 
somewhat below normal. At this speed the engine is allowed 
to run without load for a couple of hours. During this test, and 
particularly during the test under half or three-quarters load 
which follows, the tester carefully watches the bearings with 
respect to heating, and also watches for knocks or other irregu¬ 
larities in the operation of the motor. If such an irregularity 
develops it can often be localized by short-circuiting the spark 
plug first of one cylinder and then of another, thus cutting that 
cylinder out, and noting the effects. In the final test the brake 
horse power is determined, the motor being adjusted until it 
shows the amount of power which has been decided upon by the 
engineering staff as the standard to be obtained from that type 
and size of motor. 


MOTOR TESTS. 


40:} 

Testing Stands —Various methods are in use for mounting 
and driving the motors during the “running in” process. In some 
factories a portable electric motor is used which is moved from 
stand to stand and connected to the motor to be “run in” by 
means of a belt over the motor pulley and the engine flywheel. 
In other instances, the “running in” stands are arranged in 
two rows underneath a line shaft and the motors are run by 



Fig. 301.—Engine Testing Staxd. 

belts from the line shaft. Since in a factory of large size 
there are always more motors under test running under their 
own power than being “run in,” it is possible to drive the line 
shaft from the motors operating under power. Finally, the 
test stands may be arranged in sets of two, one containing a 
motor to be run in and one a motor to be tested under power, 
and the two motors connected to each other by a quickly ap- 










464 


MOTOR TESTS. 


plied coupling. The line shaft method is probably employed 
the most. Fig. 301 shows a testing stand designed to take 
engines of different sizes and to permit of quickly securing the 
engine in position and removing it. 

Defects in Valve Timing —If the motor, upon being started 
under its own power, does not operate satisfactorily, but shows 
a serious lack of power, it may be due to incorrect timing of 
the valves. The camshaft gears may have been improperly 
meshed in assembling, or the valve lifters may not be properly 
adjusted. It is now customary, at least in the case of the higher 
grade motors, to mark the points of valve opening and closing on 
the flywheel and secure a pointer to the rear cylinder extending 
on the flywheel show when the crankshaft is in the dead centre 
across to the flywheel rim. As indicated in Fig. 302, the marks 
position and when each of the different valves should begin to 
lift and should close. Therefore, if the motor fails to develop 
the usual amount of power, the tester should observe whether 



Fig. 302.—Valve Timing Marks on Flywheel. 


each of the valves actually begins to lift when the pointer is 
opposite the mark on the flywheel which indicates the opening 
position for that valve, and closes when the pointer is oppo¬ 
site the mark indicating the closing position for the valve. 
These marks are put on the flywheel after the engine is 
assembled and the proper points for the marks are in the 
first place determined as follows: Suppose that we have a 
4x5 engine with a 16 inch flywheel and that the exhaust valve 
is to begin to open 40 degrees ahead of the bottom dead centre. 
Then the distance along the circumference of the flywheel cor¬ 
responding to the advance of exhaust opening is 

40 X l6 X 3.14 _ 0*7 

- --—2 =5.58 inches. 

360 

The lengths of arc corresponding to the exhaust valve closing 
lag, the inlet valve opening lag and closing lag can be found 
in the same way. When scribing the marks upon the fly- 







MOTOR TESTS. 


465 


wheel rim, the first thing to do is to determine the two dead 
centre positions. To this end the plug or cock in the centre 
of the cylinder head is removed and a rule or rod is inserted 
through the opening and rested on top of the piston. The 
crankshaft is then turned until the piston is from i to 1^4 
inches from the top end of its stroke. Then a scratch awl 
mark is made on the flywheel rim at a point opposite the sta¬ 
tionary pointer. Next the crankshaft is turned over the dead 
centre until the piston is again the same distance from the 
top of the stroke, and another mark is made on the flywheel 
rim opposite the stationary pointer. The distance between the 
two marks on the rim is then halved, and at the point midway 
between the marks the top dead centre mark is inscribed. In a 
four cylinder motor this mark corresponds to the top dead 
centre position of two of the cranks and the bottom dead 
centre position of the other two, except if the crankshaft is 
offset. The other dead centre position is directly opposite and 
can be found in the same way. Of course, the dead centre po¬ 
sition can be found directly by turning the crankshaft over 
and noting the inserted rod rise and fall with the piston, the 
dead centre position being reached when the rod stops rising. 
But this method is not as accurate as the one previously de¬ 
scribed. The positions for other marks are then laid off 
by means of a tape scale. 

Having the marks on the flywheel, the tester can determine 
whether the valves actually begin to lift and finally close at the 
intended moments. The push rods should be so adjusted that 
when the valve is seated there is a slight clearance between 
their top and the bottom of the valve stem, so that a sheet of 
thin paper may be inserted between them. When the valve be¬ 
gins to lift the paper will be clamped between the two parts and 
it is then impossible to withdraw it. If this first occurs when 
the valve timing pointer is opposite the valve opening mark, the 
valve is properly timed as to its opening. Should it be found 
that all of the valves are out in their timing in a certain way— 
that is, either too early or too late—it is a sign that the camshaft 
gears were improperly meshed in assembling and the trouble 
can be corrected by shifting one of the gears around one tooth 
or several teeth, as the case may require. If only one or two 
of the valves are improperly timed, the only remedy short of fit¬ 
ting another cam or camshaft, consists in adjusting the push rods. 

Location of Knocks —Among the more serious defects which 
may be found in a new motor is a knock in one of the bearings 
or elsewhere. When such a defect is present it is readily de- 


466 


MOTOR TESTS. 


tected by the ear, but its exact location is a problem that is not 
so easily settled. Knocking is generally due to looseness in some 
part, most frequently in the connecting rod bearings and at the 
joint of the flywheel with the crank shaft. It requires a free 
play of only a few thousandths of an inch to produce an audible 
knock, and any slight defect in the workmanship, therefore, may 
bring on this trouble. The location of the knock is greatly 
facilitated by means of a stethoscope, as used by physicians in 
locating irregularities in the action of certain organs of the 
human body. It consists of a piece of hard rubber connected 
by rubber tubes to the observer’s ears. The instrument is used 
as shown in Fig. 303, the hard rubber end being passed over 
the engine, the noise observed increasing as the location of the 
knock is approached. 



Fig. 303.— Locating Knocks by Means of a Stethoscope. 


Compression Leaks—Another thing that has to be looked for 
in making commercial tests of engines is cylinder leaks, which 
would cause a reduction in the compression. If there is a 
material difference in the compression of the different cylinders, 
it may be detected by opening all of the compression cocks but 
one and turning the engine over by hand, or by allowing the ex¬ 
haust to discharge directly into the atmosphere, when a differ¬ 
ence in the loudness of the exhaust reports will be noticed. In 
order to determine the location of the leak, a light oil is poured 











MOTOR TESTS. 


467 


over the various joints in the cylinder wall, viz., around the 
compression cock, the spark plug, the valve caps, etc. If there is 
a leak at any one of these points it will be shown by the forma¬ 
tion of bubbles. Should no leak be found at any of these points, 
and yet the cylinder show a loss of compression, it must be due to 
either leaky valves or leaky piston rings. In such a case the 
valves should be removed and carefully examined, and if the 
examination prove that they do not seat perfectly all around, 
they should be reground first with emery dust and oil and then 
with rotten-stone and water. If the leak still continues after 
the valves have thus been ground, the trouble must be in the piston 
rings, and these should be replaced with new ones. 

When the more important defects have thus been eliminated 
the motor is subjected to the brake test. If at first it fails to 
develop the power called for further improvements are made 
in its adjustment until it finally delivers the required output, 
when it is ready for the stock room. 






APPENDIX. 


TRACTION RESISTANCE. 

When a motor-propelled vehicle travels on a level course 
the resistance to its motion consists of three main items, viz., 
the rolling resistance at the contact of the wheels with the 
ground, the friction in the rear axle and drive, and the air 
resistance. In case the vehicle travels on an up grade' there is 
another factor to be taken into account, namely, the power re¬ 
quired to lift the weight of the vehicle at a certain rate. When 
the car is running down hill there is a corresponding negative 
item. 

The road resistance proper, of course, depends very largely 
upon the character of the road surface, but it is also dependent 
upon the diameter of the wheels, the character of the tires and 
the speed at which the vehicle travels. The second item, the fric¬ 
tion of the rear axle and drive shaft bearings and gears, is com¬ 
paratively small, and since it is difficult to separate this from 
the road resistance, experimentally, it is generally included 
with it. 

On very smooth macadam pavement the traction coefficient may 
be as low as 15 pounds per 1,000, if the vehicle speed is low and 
the vehicle wheels are of large diameter and shod with rubber 
tires. It is customary, however, to figure on a traction resistance 
of 25 pounds per 1,000 for macadamized roads in fair condition 
and smooth earth roads. Denoting the total weight of the car 
with passengers by W, the total effort required to overcome 
traction resistance is evidently 

W X 2 5 founds . 

1,000 

If v be the speed of the car in miles per hour, then 

v X 5.280 _ v 
60 

is the speed in feet per minute, and the horse power required to 


469 






470 


APPENDIX. 


overcome traction resistance is 


//. /' t = X 88 v X —-— 

1,000 33. 000 


2 W V 


30.000 

The air resistance is directly proportional to the forwardly 
projected area A of the car and to the square of the speed. 
We may therefore write 

Kcl = c A v 2 founds , 

where c is a constant. The horse power required to overcome 
this resistance is 

h.p* = A±!L= c AjL 

33,000 33,000 

where R » is expressed in pounds, A in square feet and v in feet 
per minute. However, the speed is usually given in miles per 
hour, and if in the above equation v- speed in miles per hour, 

c _•_. 1 _, 1 


then the expression 

33.ooo 

from which we find that 


is approximately equal to 


100,000 


II P a = 


A ir ' 


100,000 

It is obvious that the total horse power required to propel 
a car on a level road is the sum of the two above expressions— 


rr D _ 2 IV V _ A V Z 

A I • Jl 1 • 1 | - ■ - 

30,000 1 100,000 


(98) 


Now let it be required to find a suitable gearing ratio for a 
car weighing with passengers ^,500 pounds and having a forwardly 
projected area of 12 square feet, to be fitted with a motor having 
a horse power characteristic like that shown in Fig. 239 (single 
point ignition). By means of equation (98) we can determine the 
horse power required by this car at different speeds, and plot 
the result on a co-ordinate diagram. 

If now we superpose this curve on the horse power charac¬ 
teristic of the motor so that corresponding car and engine speeds 
coincide, the point of intersection of the two curves gives the 
speed at which the car will run on the level. But the gearing 
ratio has not yet been fixed, and we therefore do not know as 
yet where to place the car horse power characteristic on the 
engine horse power diagram. 

The gearing ratio would be determined by the consideration 
that the car shall be able to climb a grade of, say, 6 per cent, on 
the high gear. The horizontal effort required to pull a car up 
a 6 per cent, grade is equal to 6 per cent, of its weight. In 













APPENDIX. 


471 


this case it is therefore 150 pounds. The traction resistance, at 
25 pounds per 1,000, amounts to 62.5 pounds. At the low speed 
at which the car will be traveling on the grade the air resistance 
will be negligible, and the total resistance to motion will there¬ 
fore be 

150 + 62.5 = 212.5 pounds. 

Now, suppose that the car has 36 inch wheels. Then there is 
a resisting force of 212.5 at a radius of iX feet, which corre- 


JYTiles Per PCour 

30 35 40 45 50 



Fig. 304. 

sponds to a torque of 318.75 pounds-feet. By reference to Fig. 
239 it will be seen that this motor (with single spark) develops 
a maximum torque of 79 pounds-feet, hence the gear ratio 
must be 

= 4 ( apflrox .) 

79 

With the motor turning at 2,000 revolutions per minute, and 
a reduction ratio from the motor shaft to the driving wheels 
of 4, the car speed would be 

36 X 3 • T 4 v, 2,000 ' 6o_ — miles fer hour. 

12 ^ 4 5,280 

























































472 


APPENDIX. 


We therefore superpose the car horse power curve and mo¬ 
tor horse power curve, making the speed scales such that 
2,000 r. p. m. = 53 m. p. h. This is done in Fig. 304, from which 
it is seen that the maximum speed of the car will be about 49 
miles per hour. Any problem in connection with the speed of 
a car can be solved in the same manner if the horse power 
characteristic of the motor is known. 

The above method of calculation enables us to determine the 
horse power required for cars of different weights and for 
different speeds desired. However, it is customary to figure 
on 1 horse power per 100 pounds of car weight for moderate 
powered pleasure cars, and 1 horse power to 60 pounds of car 
weight for high powered roadsters and touring cars. The aver¬ 
age A. L. A. M. rating of motors for trucks is given by the 
formula 

H. P. = 15 + 5 x T, 

where T is the load capacity in tons. 


Taper Pins. 

Standard taper pins for securing machinery parts together have 
a taper of Y$ inch to the foot. They are regularly made in the 
following sizes: 


Largest Range of Lengths 

Diameter. in Steps of J 4 Inch. 

Inch. Inches. Inches. 

0.156 . 34 1 34 

0.172 . y 2 

0.193 . 34 2y x 

0.219 . 34 3 

0.250 . 34 3 

0.289 . 34 3 

0.341 . 34 4 

0.409 . 1 4 

0.492 . i »4 414 

0.591 . i'A s'A 

>•7 06 . 114 6 














APPENDIX. 


473 



Plug. Plug. 

S. A. E. Standard Spark Plug Shell 

Specifications: Stock, hexagon, i Y inch across flats. Flange 
form, circular; diameter, i% inch; thickness, fir inch; face, 
annular seat Y inch wide (suitable for copper-asbestos gasket). 
Blank for thread, Y inch diameter; y 2 inch length (including 
recess). Recess adjacent to flange, Y inch wide; bottom 
diameter to equal core diameter of thread; diameter of flange to 
be not less than Y inch nor more than 57/64 inch (minimum 
radius to be determined). Thread diameter, Y inch straight 
(tolerance from precise measurement of thread diameter, + O, 
•—0.003 inch) ; pitch, 18; form, U. S. standard; cut, perfect into 
recess and cuff; both ends to be beveled to an angle of 30 
degrees with planes at right angles to the axis of the plug. 
Extension of spark plug shell below threaded portion shall be 
known as cuff. Cuff shall be cylindrical in shape and the outside 
and inside edges of the same shall be chamfered; length, 3/32 
inch; diameter, 25/32 inch. 

De Dion (European) Standard Spark Plug Shell. 

This shell has a straight thread 18 mm. in diameter with 
a pitch of 1.5 mm. The thread portion is 10 mm. long. The 
other dimensions are as shown in the accompanying cut. This 
form of plug is universally used in France and has been adopted 
as a standard in Great Britain. 


































































































474 


APPENDIX 


Toothed Gearing. 

Following are a few simple rules used in the calculation of 
the elements of toothed gearing: 


SPUR GEARS. 


Pitch diameter = 

Circular pitch = 

Addendum (height of tooth _ 
above pitch circle) 

Dedendum (depth of tooth be-_ 
low pitch circle) 

Outside diameter of gear blank = 

Bottom diameter of tooth an-_ 
nulus 


Number of teeth. 
Diametral pitch. 

_ 3-i4i6 __ 

Diametral pitch. 

_ i _ 

Diametral pitch. 

_ bl3Z _ 

Diametral pitch. 

Number of teeth + 2 
Diametral pitch. 
Number of teeth — 2.3 14 
Diametral pitch. 



In gear cutting practice the following diametrical pitches are 
used: All the integral numbers from 2 to 12, inclusive, besides 
2 ]/ 2 , all the even numbers up to and including 24. and all the 
numbers divisible by 4 up to and including 48. 

STUB TEETH, 

The common form of spur gearing has involute teeth with a 
pressure angle of \\V2 degrees. In automobile work so-called 
stub teeth are much used which are of smaller height and have 
a greater pressure angle, Stub teeth are generally designated 
by two pitch numbers, as, for instance, 6-8 pitch. This means 
that for a gear of a given number of teeth the pitch diameter is 
figured as for a 6 pitch regular tooth, and the addendum and 
dedendum are figured as for an 8 pitch regular tooth. Since 












APPENDIX. 


475 


the height of the stub tooth is less, while its thickness is the 
same as that of the regular tooth, it is considerably stronger 
than the latter, the proportional increase in strength being 
especially great in the case of gears with small numbers of 
teeth. Owing to the increased pressure angle the pressure on 
the bearings is somewhat greater with the stub tooth, but this 
disadvantage is slight and is more than balanced by the ad¬ 
vantage that there is less sliding at the contacts in the stub 
tooth. The following rules hold for stub tooth gears cut ac¬ 
cording to the double pitch system: 


Pitch diameter 
Circular pitch 


Number of teeth, 
smaller pitch number. 

_ 3-i4i6 _ 

smaller pitch number. 


Addendum 

Dedendum 
Outside diameter of 

gear blank = pitch diameter 

Bottom diameter of tooth an¬ 
nulus 


_1_ 

larger pitch number. 

_PI 57 _ 

larger pitch number. 

2 

larger pitch number. 
Pitch diameter — 2.314 
larger pitch number. 


BEVEL GEARS. 


Bevel gears, as a rule, are used only for transmitting motion 
between shafts at right angles, though they may be used for 
transmission between shafts at any angle in the same 
plane. All elements of the tooth surfaces converge toward 
the point of intersection of the axes of the gears. The teeth 
therefore decrease in size from the outer toward the inner end. 
The calculation of the pitch diameters, outside diameters and 
bottom diameters is substantially the same as in the case of 
spur gears, the calculations being made for the larger end of 
each gear. The use of the formulae may be illustrated by an 
example. 

Suppose that a pair of bevel gears are to be designed to 
work at right angles and to give a ratio of 1 to 2. Then the 
gear must have twice the number of teeth as the pinion, of the 
same pitch. Suppose that the amount of power to be trans¬ 
mitted is such that gears with 18 and 36 teeth of 6 pitch with 
1 inch face appear to give the proper strength. Then the largest 
pitch diameters will be 3 and 6 inches respectively. These diam¬ 
eters can be laid off as shown in Fig. 308. The addenda for 








476 


APPENDIX. 


both gears are 1-6 inch and the dedenda ^ =0.193 inch. 

These distances are laid off on lines perpendicular to lines 
drawn from the intersection of the axes of the gears to the 
ends of the largest pitch diameters. 

In turning up the blanks it is necessary to know the face 
angle. This can be found either by graphical construction or 
by calculation. The method of calculation (taking the pinion as 
an example) is as follows: 

1 • 5 

The tangent of one half the pitch angle is =0.5. There¬ 
fore, one-half the pitch is 26°34' and the whole pitch angle 55 0 
8'. The face angle is larger than the pitch angle by twice the 
angle subtended at the point of intersection of the axes by the 


-Pace Angle 
—Pilch Angle 
Bo Horn Angle 


Pitch £1 a meter 
Outside Diameter 



Fig. 308. —Bevel Gear Lay-out. 


tooth addendum. This angle may be found as follows: The 
distance from the point of intersection to one end of the pitch 
diameter is 

x/32 X 1.5 2 = 3-354 

Now the addendum divided by this distance gives the tangent 
of the angle subtended by the addendum at the point of inter¬ 
section. 

o. 167 

- - =0.05 ap$r. 

3-354 






















































APPENDIX. 


477 


From a table of tangents we find that an angle of 2 degrees 
52 minutes has a tangent of 0.05. Hence the face angle is 

55 0 8' X 2 (2 0 52O = 6o° 52' 

The bottom angle can be found in a similar manner. 

The sides of a bevel gear are turned up at right angles to the 
pitch line. For turning up the blank, the dimensions required 
are the face, pitch angle, face angle and outside diameter. The 
latter may be obtained by means of the equation: 

Outside diam. = pitch diam. + 2 x addendum x cos of pitch angle. 

HELICAL GEARS. 

In gears in which the teeth are not parallel with the axis of 
rotation a distinction is made between the normal pitch and 
the apparent pitch. The apparent pitch is found by dividing 
the number of teeth by the pitch diameter. The normal pitch 



is obtained by dividing the apparent pitch by the cosine of the 
angle of spiral. The angle of spiral is the angle made by a 
tangent to an element of the tooth surface in a helical gear with 
an intersecting line parallel with the axis of the gear. The 
spiral advance is the distance in the direction of the axis of 
the gear in which the elements of the tooth surface describe 
a complete circle. The following rules apply: 

_ Number of teeth. _ 

Cosine of angle of spiral - N orma i pitch x pitch diameter. 

3.1416 x pitch diameter 

Spiral advance - tangent of angle of spiral. 

Outside diameter = pitch diameter+ 


normal pitch. 





















478 


APPENDIX. 


The addendum and dedendum are the same as in a spur gear 
of the same pitch as the normal pitch of the helical gear. 

Helical gears may be cut so as to properly mesh together 
when mounted on shafts making any desired angle with each 
other. The most common cases are where the shafts are parallel 
and where they are at right angles with each other, respectively. 
When the shafts are parallel the angles of spiral must be the 
same for both pinion and gear, but one gear must be left handed 
and the other right handed. When the two shafts are at right 
angles with each other the sum of the angles of spiral of the two 
gears must be 90 degrees, and both helices must be either 
right handed or left handed. The only object in using helical 
gears when the shafts are parallel is to make the gears more 
nearly quiet. In such cases, in order to keep down the end 
thrust, the angle of spiral is made not to exceed 30 degrees. 
The direction of the end thrust changes with the direction of 
the helices. When the shafts are at right angles the angles of 
spiral of both gears are generally not far from 45 degrees. 
The ratio between the speeds of revolution of two helical gears 
in mesh with each other is the reciprocal of the ratio of their 
respective numbers of teeth. 

METRIC GEAR STANDARDS. 

In countries employing the metric system of weights and 
measures gear calculations are based upon the module, which 
corresponds to the reciprocal of our diametral pitch. All di¬ 
mensions are in millimeters and the rules for spur gears are 
as follows: 

Pitch Diameter 
Number of teeth. 

= 3.1416 x module. 

- Module. 

= 1.157 x module. 

Outside diameter of gear blank = Pitch diameter + (2 x module). 

The standard modules progress in steps of from 1 to 5P2. 
then in steps of l / 2 to 7, and then in steps of 1. 


Module 

Circumferential module 

Addendum 

Dedendum 




APPENDIX. 


479 


S. A. E. Standard Screws and Nuts. 



B—Refers to all nuts and screw heads. 
D—Diameter of screw, 
d—Diameter of cotter pin. 

Dxi.s—Length of threaded portion. 

P—Pitch of thread. 

P 

-Flat top. 

8 


D 

p 

A 

Ax 

B 

C 

E 

H 

I 

K 

d 


28 

$2 


t 7 s 

$2 

A 


$2 

tV 

T*8 


24 

u 

hi 

34 

£2 


hi 

3 ^ 

t's 

ik 

Ys 

24 

tt 

H 

tS 

Ys 

Ys 

#2 

34 

^2 

sk 

r 7 3 

20 

n 

Ys 

Ys 

Ys 

Ys 

n 

34 

T$2 

Su2 

34 

20 

& 

T?S 

Y 

& 

Ys 

Ys 

34 

$2 

$2 

* 

18 

30/6 4 

u 

Ys 

ik 

& 

n 

34 

$2 

Ys 

Ys 

18 

§1 

35/64 

It 

Ya 

& 

** 

34 

$2 

Ys 

Ih 

16 

49/64 

i* 

1 

Ya 

£ 2 

33/64 

34 

$2 

Ys 

Ya 

16 

it 


I T*B 

Ya 

/t '2 


Ys 

$2 

Ys 

Vs 

14 

§1 

49/64 

1 34 

Ya 

'h 

u 

34 

$2 

Ys 

i 

14 

1 

Vs 

ilk 

Ya 

$2 

Ya 

34 

$2 

34 

i 34 

12 

1^2 

63/64 

1 Ys 


3 h 

ll 

3 ? 2 

$2 

hi 

i 34 

12 

1 Va 

1^2 

ill 


3^2 

11 

3*2 

$2 

hi 

iYs 

12 

ill 

itt 

2 

Ys 

34 

I rfe 

34 


hi 

i 34 

12 

i 34 



Ys 

34 

134 

34 


hi 


Dimensions—All dimensions in inches. 

Finish—All heads and nuts to be semi-finish. 

Material—For all screws and nuts—steel. 

Screws are to be left soft. Screw heads are to be left soft. 
The plain nuts are to be left soft. The castle nuts are to be case 
hardened. 

















































480 


APPENDIX. 


It is assumed that where screws are to be used in soft material, 
such as cast iron, brass, bronze or aluminum, the United States 
standard pitches will be used. 

Tolerance—The body diameter of the screws shall be one- 
thousandth (o.ooi) inch less than the nominal diameter, with 
a plus tolerance of zero and a minus tolerance of two-thousandths 
(0.002) inch. 

The nuts shall be a good fit without perceptible shake. 

Clearance between tops of threads and bottoms of threads— 
The tap shall be between two-thousandths (.002) inch and three- 
thousandths (.003) inch large. 

Drill Sizes, Cross Sectional Areas and Tensile Strengths of 

S. A. E. Standard Screws. 


Areas.- 

-Tensile Strengths.- 

Bottom 

At 20,000 At 25,000 At 30,000 


Nominal Drill 

Full 

of 

Lbs. Per 

Lbs. Per 

Lbs. Per 

Diameter. Size. 

Bolt. 

Thread. 

Sq. Inch. 

Sq. Inch. 

Sq. Inch. 

Ya 

$2 

O.O49I 

O.O325 

651 

814 

977 

fs 

ti 

O.0767 

0.0525 

1,050 

1,312 

i ,575 

Vs 

H 

O.IIO4 

O.0808 

1,617 

2,021 

2,426 

tV 

Vs 

0.1503 

O.II32 

2,264 

2,830 

3,396 

K 

T 7 ? 

O.I963 

O.I486 

2,972 

3,726 

4,459 


V2 

O.2485 

0.1888 

3,777 

4,722 

5,666 

H 

9/l6 

O.3068 

O.240O 

4,800 

6,000 

7,200 

tt 

39^4 

0.3712 

0.2888 

5,776 

7,220 

8,664 

H 

43/64 

O.4418 

0.3514 

7,028 

8,785 

10,452 

Vs 

§5 

O.6013 

O.4816 

9,633 

12,141 

14,449 

1 

§1 

O.7854 

O.6463 

12,926 

16,158 

19,389 

iVs 

Is 1 ? 

0.994 

0.799 

15,980 

19,975 

23,970 


I& 

1.2272 

1.009 

20,180 

25,225 

30,270 

iH 

I hi 

I.4849 

1.243 

24,860 

31,075 

37,290 

l l /2 

iH 

I.7671 

1.503 

30,060 

37,575 ' 

45,090 



Standard Metric Threads. 


The following table gives the pitches of 

screws of different 

diameters cut 

with standard metric threads 



Screw Diam. 

Pitch. 


Screw Diam. 

Pitch. 


Mm. 

Mm. 


Mm. 


Mm. 


6 

I 


16 


2 


7 

I 


18 


2-5 


8 

1-25 

20 


2-5 


9 

1.25 

22 


2.5 


10 

1-5 


24 


3 


11 

1.5 


27 


3 


12 

i -75 

30 


3-5 


14 

2 


33 


3-5 







APPENDIX. 


481 


Drill Sizes, Cross Sectional Areas and Tensile Strengths of 

U. S. Standard Screws. 


Nominal 


Drill 

-Areas.- 

Full Bottom 

Tensile Strengths. 
At 15,000 At 20,000 
Lbs. Per Lbs. Per 

Diameter. 

Pitch. 

Size. 

Bolt. 

of Thread. 

Sq. In. 

Sq. In. 


20 

3 /l 6 

O.O49I 

0.0269 

403 

538 

& 

18 

1/4 

O.0767 

O.0454 

68l 

908 

Vs 

l6 

5/16 

O.TIO4 

O.0678 

1,017 

1,356 

T^T 

14 

23/64 

0.1503 

O.0934 

1,401 

1,868 


12 

13/32 

O.I963 

O.II 55 

L 732 

2,310 

Vs 

II 

17/32 

O.3068 

0.2018 

3,027 

4,036 

3/4 

10 

41/64 

O.4418 

0.302 

4,530 

6,040 

7 A 

9 

3/4 

0.6013 

O.4193 

6,290 

8,386 

1 

8 

55/64 

O.7854 

0 . 55 H 

8,266 

11,022 


S. A. E. Standard Magneto Dimensions for Four and 

Six Cylinder Engines. 


Mm. Inches 

Shaft height. 45. 1.771 

Distance from center of front base-plate 

holes to large end of shaft taper. 53. 2.086 

Distance from center of front base-plate 

holes to center of rear base-plate holes.. 50. 1.968 

Distance between centers of base-plate holes 

left to right. 50: 1.968 

Large diameter of taper. 15. 0.590 

Small diameter of taper. 12. 0.472 

Length of taper. 15. 0.590 

Taper 1:5 (included angle) n deg. 30 min. 
approx. Woodruff key No. 3. 

Height of magneto space. 203. 8.000 

Length of magneto space. 254. 10.000 

Width of magneto space. 127. 5.000 

Plain hole timing lever. 6.35 0.25 

Tapped hole timing lever, *4 inch, 28 P., 

S. A. E. 


Base-plate holes— Y> in. 16 P., U. S. S. 

Thread for end of magneto shaft Y in., 

16 P., U. S. S., length of thread 0.5905 in. 

Advance lever radius. 2.125 in. 

















482 


APPENDIX. 

Silent Chains. 


Let P = pitch of chain and 

A^ = number of teeth in sprocket, 

Then 


F 

Pitch diameter = ^ inches 

sin - 

N 



Let Z.- number of links in chain; 

D - pitch diameter of sprocket wheel; 
d = pitch diameter of sprocket pinion; 

N = number of teeth in wheel; 
n = number of teeth in pinion; 

L — distance between axes of wheel and pinion; 
Then 


where 


z = -f (/V—«) A + 


180 


2 D cos (3 ' 
F ’ 


2 L 

The above formula can be used for determining the approxi¬ 
mate number of links required for a certain centre distance and 
sprockets with certain numbers of teeth. But since the result 
is almost sure to involve a fraction it cannot be used directly. 
















APPENDIX. 


483 


POSSIBLE CENTRE DISTANCES FOR “ONE TO TWO” PAIRS OF 

SILENT CHAIN SPROCKETS. 

(Chains running tight on sprockets.) 


No. 

No. 




Pitch.- 


of Teeth 

of Teeth 

No. 

0.315 

0.4 

0.5 

0.625 



of Links 





x 11 Vj v c1 r • 





13 

26 

33 

2.02$ 

2.565 

3.206 

4.008 

13 

26 

34 

2.186 

2.776 

3-470 

4-337 

14 

28 

36 

2.252 

2.859 

3.574 

4.468 

15 

30 

38 

h) 

O0 

H 

00 

2.934 

3.680 

4.600 

13 

26 

35 

2.350 

2.984 

3-730 

4.663 

14 

28 

37 

2.417 

3-069 

3.836 

4.796 

15 

30 

39 

2.483 

3-154 

3-942 

4.928 

13 

26 

36 

2.514 

3.192 

3-990 

4.987 

16 

32 

4 i 

2.550 

3-238 

4.048 

5.060 

14 

28 

39 

2.745 

3-279 

4.098 

5-122 

17 

34 

43 

2.616 

3-322 

4.152 

5.190 

15 

30 

40 

2.649 

3-363 

4.204 

5-255 

13 

26 

37 

2.677 

3-399 

4.248 

5-310 

16 

32 

42 

2.715 

3-448 

4.310 

5.388 

14 

28 

39 

2.745 

3-486 

4-357 

5-447 

1 7 

34 

44 

2.782 

3-533 

4.416 

5.520 

IS 

30 

4 i 

2.812 

3-571 

4.464 

5 . 58 o 

13 

26 

38 

2.838 

3.604 

4.505 

5-632 

18 

36 

46 

2.848 

3.616 

4.520 

5-650 

16 

32 

43 

2.880 

3-658 

4-572 

5 . 7 I 5 

14 

28 

40 

2.907 

3.692 

4.615 

5.769 

19 

38 

48 

2.914 

3.700 

4.625 

5.782 

. 17 

34 

45 

2.947 

3-742 

4.678 

5.848 

IS 

30 

42 

2.925 

3-778 

4-723 

5-904 

13 

26 

39 

3.000 

3-8io 

4.762 

5-953 

18 

36 

47 

3.013 

3.828 

4.783 

5-979 

16 

32 

44 

3-044 

3.866 

4.832 

-6.040 

14 

28 

41 

3-069 

3-898 

4.872 

6.090 

19 

38 

49 

3.080 

3.901 

4.889 

6.112 

17 

34 

46 

3 - no 

3-950 

4-937 

6.172 

15 

30 

43 

3.037 

3-984 

4.980 

6.226 

20 

40 

5 i 

3.146 

3-994 

4-993 

6.242 

13 

26 

40 

3- i 6 i 

4.014 

5.018 

6.273 

18 

36 

48 

3-178 

4-035 

5-044 

6.305 

16 

32 

45 

3.207 

4.072 

5.090 

6.363 

14 

28 

42 

3-231 

4-103 

5.129 

6.411 

19 

38 

50 

3.246 

4.122 

5-152 

6.440 

1 7 

34 

47 

3-274 

4.158 

5-197 

6-497 

IS 

30 

44 

3-299 

4.220 

5-237 

6.547 

20 

40 

52 

3-312 

4.201 

5.257 

6.571 

13 

26 

41 

3.322 

4.218 

5-273 

6.592 

18 

36 

49 

3-342 

4-243 

5-304 

6.630 

16 

32 

46 

3.369 

4.278 

5-348 

6.685 

14 

28 

43 

3-392 

4-307 

5-384 

6.731 

19 

38 

5 i 

3.410 

4-330 

5 . 4 I 3 

6.766 

17 

34 

48 

3.438 

4-366 

5-457 

6.821 

16 

30 

45 

3 . 46 l 

4-395 

5-494 

6.867 

20 

40 

53 

3-477 

4-415 

5.519 

6.899 

1 3 

26 

42 

3-483 

4.422 

5.528 

6.910 

18 

36 

50 

3.505 

4.451 

5.564 

6.955 

16 

32 

47 

3.531 

4.484 

5-605 

7.006 






484 


APPENDIX. 


The correct centre distance for the nearest full number of links 
may then be found by means of the following equation : 


L = 


P [z —— _ — — {N — n) -£-] 

22 180J 


2 COS ft 

The accompanying table gives all the possible centre distances 
for “one to two” sets of silent chain sprockets over a certain 
range of sprocket sizes and link numbers for the four smallest 
pitches in practical use. This range is believed to include all 
possible cases of automobile camshaft drives. 

The outside diameter of the sprocket is different for the dif¬ 
ferent makes of chains. In some makes it is slightly greater than 
the pitch diameter and in others slightly less. This dimension, 
therefore, can only be determined after the make of chain to be 
used has been settled upon. 



Fig. 312.—Centre Distance Diagram. 


















APPENDIX. 485 

Dimensions of Cylinder Ring Blanks (Eccentric Rings). 



Fig. 313. —Piston Ring Blank. 






Width 

Width 



Outside 

Inside 

Eccen¬ 

of Squares 

of 45 0 



Diameter. 

Diameter. 

tricity. 

Slot. 

Diagonal Slot. 

Width. 

Bore. 

(Do) 

(Di) 

(E) 

(S) 

(S45) 

(W) 

3 

3.105 

2.919 

.022 

0.270 

0.190 

.150 

3 X 

3-234 

3.041 

.028 

0.281 

0.197 

.156 

3 l A 

3-364 

3.162 

.029 

0.292 

0.204 

. 163 

3X 

3-493 

3-284 

.031 

0.304 

0.213 

. 169 

3 l A 

3.622 

3-405 

.032 

0.315 

0.220 

•175 

3H 

3-752 

3.527 

.033 

0.326 

0.228 

.181 

3H 

3-88i 

3-649 

•034 

0-337 

0.236 

r-x 

00 

• 

3 X 

4.011 

3-770 

.035 

0-349 

0.244 

.194 

4 

4.14 

00 

vO 

.036 

0.360 

0.252 

.200 

4H 

4.269 

4.014 

• 0 37 

0-371 

0.260 

. 206 

. 4 A 

4-399 

4.135 

• 039 

0.382 

0.267 

.212 

4 X 

4.528 

4.257 

.040 

0-394 

0.276 

.219 

4*A 

4-657 

4.378 

.041 

0.405 

0.283 

.225 

4% 

4.787 

4-500 

.042 

0.416 

0.291 

.231 

4V4 

4.916 

4.622 

.043 

0.427 

0.299 

.237 

4 x 

S.046 

4-743 

.044 

0-439 

0.307 

.244 

5 

5-175 

4-865 

.045 

0.450 

0.315 

.250 

sX 

5-304 

4.987 

• 047 

0.461 

0.322 

.256 

sX 

5-434 

5.108 

.048 

0.472 

0.330 

.262 

sX 

5-563 

5-230 

.049 

0.484 

o .339 

.269 

s'A 

5-693 

5-351 

.050 

0-495 

0.346 

■ 275 










486 


APPENDIX. 


Valve Spring Table. 

In the following table D denotes the mean diameter of the coil 
(from centre to centre of wire), w’hich is equal to the outside diam¬ 
eter minus the diameter of the wire; d, the diameter of the wire 
IV, the maximum safe pressure a spring of the particular diam¬ 
eter will sustain, and F the deflection of one coil under a 
pressure of too pounds. It will be noticed that three different 
values are given for F for each size of wire and diameter of 
coil; these correspond to coefficients of torsional elasticity of 
10,000,000, 12,000,000 and 14,000,000, respectively. The maximum 
safe pressure is calculated on the basis of a stress of 50,000 
pounds per square inch. The table covers every size of wire 
in the Birmingham gauge from No. 11 to No. 6, inclusive. 

An example will show the use of the table. Let us suppose 
that we have a spring wound from No. 8 wire, with a mean 
diameter of coil of 1 inch. It will be seen that the maximum 
safe pressure of such a spring is 88.2 pounds. Assuming the 
coefficient of torsional elasticity to be 12,000,000, the deflection 
of each coil under a load of 100 pounds will be 0.0899 inch. 
Therefore, if there are ten coils in the spring and the maximum 
load is 80 pounds, the total compression of the spring from 
the free state will be 


10 X 80 X o. 0899 
100 


= 0 719 inch 



D X 

X 

1 

1 X 

I Va 




N_^ 

W 45.2 

38.8 

33-9 

30.2 

27.1 




0 d 

. / .1628 

•2585 

•3858 

• 5312 

•7538 




II £ 

F ■) .1357 

.215 

.3215 

.4427 

.6282 




*0 

( .1162 

.1846 

.2756 

• 3795 

•5385 





D H 

Vi 

1 

1 X 


iX 




W 63 

54 

47-2 

42 

37-8 

34-4 



0 • 

( .1047 

.16 62- 

.2481 

•3417 

.4847 

.6436 



"1 

F ) .0872 

• 1385 

.2068 

.2847 

• 4039 

■5363 



*0 

/ -0748 

.1187 

•1773 

.2498 

.3462 

•4598 




D y 4 

Vi 

1 

1 Vi 

1 X 

1 X 




W 84.9 

72.7 

63.6 

56.6 

50.9 

46.3 

42.4 


• • 

O O 

( .0703 

.1117 

.1667 

.2296 

•3257 

•4335 

.5628 


11 £ 

K J .0586 

.0931 

.1389 

.1913' 

.2714 

.3612 

.4690 



( .0502 

.0798 

.1191 

.1640 

.2326 

.3096 

.4020 


IT) 

D X 

Vi 

1 

I X 

1 Va 

lX 

1 

1 Yi 

'S « 

W 117.6 

100.8 

88.2 

78.4 

70.6 

64.1 

58.8 

54-3 

0 0 

/ .0396 

.0723 

.1080 

.1486 

.2108 

.2806 

•3643 

•4632 

II z 

F .0379 

.0603 

.0899 

.1238 

•1757 

.2338 

.3036 

.3860 


( .0283 

•0517 

•0771. 

.1062 

.1506 

.2004 

.2602 

•3309 







APPENDIX. 


-187 



D I 




ij 4 

1 H 

iy A 

1 yi 

00 rx 

W 128.4 

.118.6 

109.5 

101.1 

93-9 

87.6 

82.1 

77-3 

d 

( .0762 

.1049 

.1488 

.1981 

-2572 

.3270 

.4085 

.5024 

11 * 

F •) .0635 

.0874 

.1240 

.1651 

.2143 

.2725 

.3404 

.4186 


f -0544 

.0610 

.1063 

.1415 

•1837 

.2336 

.2918 

• 3589 


D 1% 

tH 

i/j 

1 H 

jM 

1 H 

2 

2 %' 

O 'O 

: W 128.4 

118.6 

109.5 

101.1 

93-9 

87.6 

82.1 

77 - 3 - 

• d 

l .0920 

.1225 

.1590 

.2021 

•2525 

•3105 

•3769 

.4520 

II £ 

•>3 

F -j .0766 

.1020 

.1325 

.1684 

.2104 

.2588 

•3141 

.3767 


f -0657 

.0875 

.1136 

.1444 

.1803 

.2218 

.2692 

.3229 


Woodruff Keys. 

Woodruff keys are used very extensively in automobile con¬ 
struction, in the motor as well as in other parts of the chassis. 
The following rules may be used in selecting the proper size of 
keys of this type: 

Let D- depth in shaft; 

A = diameter of shaft; 

T = thickness of key; 
then, for most ordinary purposes, 



4 

Where the key has to perform exceptionally severe service, 
as in the case of a flywheel, a propeller shaft gear, etc., it is ' 
advisable to calculate it by means of the following equation: 

F 

Z e X T= - 

*S*9 

where 

Lc = effective length of key; 

F = maximum force at radius of key; 

S g = allowable shearing stress. 

In some cases where the small thickness of the hub shell does 
not admit of the use of a single key of adequate size two smaller 
keys may be used arranged end to end. However, if the hub 
is not long enough for this, the two keys may be placed at 90 
degrees with each other. The former arrangement is the least 
expensive, but the latter fits by far the greater number of cases. 
The depth of the key in the hub, etc., is made one-half the thick¬ 
ness of the key, as with common square keys. The following 
table of Woodruff key dimensions is taken from an article on 
the subject by George E. Goddard in The Horseless Age, of 
December 2, 1908: 






488 


APPENDIX. 



Fig. 314. —Woodruff Key. 


Dimensions of Woodruff Keys. 


No. 

of 

Key 

Thick¬ 

ness 

T. 

Approxi¬ 

mate 

Length 

L. 

Depth 
in Shaft 
D. 

V 

Depth 
in Piece 

d. 

Whole 
Width 
of Key 
W. 

Radius 

of 

Key 

R. 

Com¬ 
plement Width 
of Radius of Flat 
C. F. 

1 

1 

16 

X 

.175 

.030 

13 

64 

X 

3 

64 

2 

• 

~h 

X. 

.195 

.050 

13 

64 

X 

3 

64 

3 

Vs 

X 

.145 

.060 

13 

64 

X 

3 

64 

4 

T2 

Vs 

.210 

.050 

X 

fe 

JL 

16 

5 

Vs 

Vs 

.190 

.060 

X 

5 

16 

1 

16 

6 


Vs 

.175 

.080 

X 

5 

16 

1 

16 

7 

Vs 

V 

.255 

.060 

5 

16 

X 

_3_ 

16 

8 

5 

32 

X 

.235 

.080 

5 

16 

Vs 

1 

T6 

9 

3 

16 

X 

.220 

.095 

5 

16 

Vs 

1 

16 

10 

5 

32 

Vs 

.300 

.080 

X 

7 

16 

1 

16 

11 

A 

Vs 

.285 

.095 

Vs 

7 

T6 

1 

16 

12 

7 

32 

Vs 

.270 

.110 

Vs 

T6 

1 

16 

A 

X 

Vs 

.255 

.125 

Vs 

T6 

1 

16 

13 

_a_ 

16 

1 

.345 

.095 

1% 

X 

1 

16 

14 

7 

32 

1 

.360 

.080 

16 

X 

1 

16 

15 

X 

1 

.315 

.125 

lV 

X 

A 

B 

5 

16 

1 

.285 

.155 

7 

16 

X 

it 

16 

3 

16 

IVs 

.390 

.095 

31 

64 

9 

T6 

5 

6T 

17 

7 

32 

IX 

.315 

.110 

31 

64 

9 

16 

5 

64 

18 

X 

IVs 

.365 

.125 

3JL 

64 

9 

16 

5 

64 

C 

5 

16 

IVs 

.335 

.155 

31 

64 

TS 

5 

64 

19 

3 

16 

IX 

.460 

.095 

35 

64 

Vs 

5 

64 

20 

7 

32 

IX 

.440 

.110 

35 

64 

Vs 

5 

64 

21 

X 

ix 

.425 

.125 

35 

64 

Vs 

5 

64 

D 

5 

16 

ix 

.395 

.160 

35 

64 

Vs 

5 

64 

E 

X 

IX 

.365 

.185 

35 

64 

Vs 

^ 1 

22 

X 

IVs 

.475 

.125 

19 

3T 

11 

16 

T2 

23 

~16 

IVs 

.440 

.155 

19 

3T 

li 

16 

T2 

F 

Vs 

IVs 

.410 

.185 

19 

1T2 

11 

16 

ih 

24 

X 

IX 

.520 

.125 

41 

61 

X 

6T 














APPENDIX. 

Dimensions of Woodruff Keys—Continued. 


489 


No. 


Approxi¬ 



Whole 

Radius 

Com¬ 


Thick¬ 

mate 

Depth 

Depth 

Width 

of 

plement 

Width 

of 

Key 

ness 

T. 

Length 

L. 

in Shaft 
D. 

in Piece 
d. 

of Key 
W. 

Key 

R. 

of Radius of Fla 
C. F. 

25 

5 

16 

IX 

.490 

.155 

41 

64 

X 

7 

64 


G 

Vs 

ix 

.460 

.185 

tt 

X 

7 

64 


26 

~16 

1 23 

1 32 

.440 

.095 

1 7 

32 

11 

16 

H 

T2 

27 

X 

1 23 

1 3“? 

.410 

.125 

1 7 

32 

11 

16 

17 

32 

T2 

28 

5 

16 

1 23 

1 32 

.380 

.155 

1 7 
32 

11 

16 

1 7 

32 


29 

X 

1 23 

1 32 

.350 

.185 

17 

32 

11 

16 

u 

T2 

R 

X 

0 5 
"T5 

.630 

.125 


IVs 

Vs 

Vs 

S 

5 

16 

9_5_ 

"16 

.600 

.155 

X 

IVs 

Vs 

Vs 

T 

X 

9-5- 

"16 

.565 

.185 

X 

IVs 

Vs 

Vs 

U 

T6 

9JL 

"16 

.535 

.220 


IVs 

Vs 

Vs 

V 

X 

2& 

.505 

.250 

% 

IVs 

Vs 

Vs 

30 

Vs 

2^ 

.755 

.185 

15 

T6 

ix 

13 

16 

A 

31 

T6 

2Vs 

.720 

.220 

15 

16 

ix 

13 

16 

3 

16 

32 

X 

2% 

.690 

.250 

15 

16 

ix 

13 

16 

3 

16 

33 

9 

16 

2V 

.660 

.280 

2_5 

16 

ix 

13 

16 

3 

16 

34 

Vs 

2% 

.630 

.315 

15 

16 

ix 

13 

16 

3 

16 

35 

11 

16 

2V 8 

.600 

.345 

15 

16 

IX 

13 

T6 

tV 

36 

X 

2V 8 

.565 

.375 

15 

16 

ix 

13 

16 

1JS 


Automobile Engine Light Lubricating Oil. 

S. A. E. Specification. 

Oil for this purpose must be a pure mineral oil, no addition 
■or adulterant of any kind being permitted. 

The following characteristics are desired: 

Specific gravity.28° to 32 0 Beaume 

Flash point, not less than.400° F. 

Fire test, not less than.450° F. 

Viscosity at ioo 0 F., Saybolt viscosimeter, not 

over .300 seconds 

Viscosity at 210° F., Saybolt viscosimeter... .40 to 50 seconds 
Viscosity at 210 0 F., Tagliabue viscosimeter. .60 to 65 seconds 
Carbon residue, not over.0.50 per cent. 








490 


APPENDIX. 


S. A. E. Bearing Metal Specifications. 

Babbitt Metal. 


Tin . 84.00% 

Antimony . 9.00% 

Copper . 7.00% 

A variation of 1 per cent, either way will be permissible in the 
tin, and .5 per cent, either way will be permissible in the anti¬ 
mony and copper. The use of other than virgin metals is pro¬ 
hibited. No impurity will be permitted other than lead, and that 
not in excess of 0.25 per cent. 

NOTE: This grade of babbitt contains an unusually large 
amount of copper, and is used for the connecting rod linings of 
motor bearings and other bearings subjected to severe service. 


White Brass. 


Copper . 3.00 to 6.00% 

Tin, not less than. 65.00% 

Zinc . 28.00 to 30.00% 

Metal with more than 0.25 per cent, impurities may be rejected. 
NOTE: This alloy gives good results in automobile engines, 
but provision should be made to have it generously lubricated. 


Phosphor Bronze. 


Copper .., 

Tin . 

Lead . 

Phosphorus 


80.00% 
10.00% 
10.00% 
0.05 to 0.25% 


Impurities in excess of 0.25 per cent, will not be permitted. 

NOTE: This is a metal similar to that specified by many rail¬ 
roads for various purposes. It is an excellent composition where 
good anti-friction qualities are desired, standing up exceedingly 
well under heavy loads and severe usage. In automobile con¬ 
struction it should be used only against hardened steel. 














APPENDIX. 


491 


S. A. E. Standards for Non=Ferrous Metal Tubing. 

Tubing can be furnished in copper, and the commercial alloys 
of copper and zinc, such as high brass, bronze, phosphor bronze 
and tobin bronze. 

Composition .—As specified to meet the requirements of use. 

Temper .—As specified in the order; may be hard, half hard 
or annealed. If annealed, the tubing may be soft or light 
annealed. 

Size Variation .—On inside and outside diameter and the thick¬ 
ness of the walls, as follows: 


Outside and Inside Dimensions. 

Up to J 4 in. inclusive.0.002 in. over or under 


Over 

Yt 

in. to 

and 

including 

H 

in. 


in. 

over 

or 

under 

Over 

H 

in. to 

and 

including 

1 

in. 


in. 

over 

or 

under 

Over 

1 

in. to 

and 

including 


in. 

.0.0035 

in. 

over 

or 

under 

Over 

1 V\ 

in. to 

and 

including 


in. 


in. 

over 

or 

under 

Over 

i'A 

in. to 

and including 

iH 

in. 


in. 

over 

or 

under 

Over 


in. to 

and 

including 

2 

in. 


in. 

over 

or 

under 


Over 2 in. l /\ of 1 per cent, over or under 

No combination of variations on the same tube shall make the 
thickness of the wall vary from the nominal by more than the 
following amounts : 


Thickness of Wall. 


Up to and including 1/64 in. 

Over 1/64 in. to and including 1/32 
Over 1/32 in. to and including 1/16 
Over 1/16 in. to and including % 
Over in. to and including J 4 

Over in. to and including 5/16 

Over 5/16 in. to and including 


.0.001 in. over or under 

in.0.002 in. over or under 

in.0.003 in. over or under 

in.0.005 in. over or under 

in.0.008 in. over or under 

in.0.0125 in. over or under 

in.0.015 in. over or under 


Special Limits .—On all stock where the above commercial 
variations are not permissible limits shall be specified in the order. 

















492 


APPENDIX. 


MetrioAmerican Units Conversion Tables. 

American to Metric. 

Inches X 2.54 = centimeters. 

Feet X 0.3048 = meters. 

Yards X 0.9144 = meters. 

Miles X 1.609 = kilometers. 

Square inches X 6.452 = square centimeters. 

Square feet X 0.0929 = square meters. 

Cubic inches X 16.387 = cubic centimeters. 

Cubic feet X 0.0283 = cubic meters. 

Pounds X 0.4536 = kilograms. 

U. S. gallons X 37854 = litres. 

British thermal units X 0.252 = calories. 

Horse powers X 1.0138 = chevoux (metric horse powers). 
Foot-pounds (work) X 0.1383 = kilogrammeters. 

Pounds feet (torque) X 0.1383 = meter-kilograms. 

Feet per minute X 0.00508 — meters per second. 

Miles per hour X 1.609 — kilometers per hour. 

Pounds per square inch X 0.0703 = kilograms per square 
•centimeter. 

Pounds per gallon X 0.1198 = kilograms per litre. 

British thermal units per pound X 0.5556 = calories per 
kilogram. 

British thermal units per gallon X 0.06658 = calories per litre. 
Pounds per horse power hour X 4474 = grams per cheval- 
heure. 

Miles per gallon X 0.425 = kilometers per liter. 

Horse power X 746 = watts. 

U. S. gallons X 0.833 — Imperial gallons. 

(Fahrenheit temp. — 32) X 0.555 — centigrade temp. 

Metric to American. 

Millimeters X 0.03937 = inches. 

Meters X 39.37 = inches. 

Meters X 3.281 = feet. 

Meters X 1.094 = yards. 

Kilometers X 0.6214 = miles. 

Square centimeters X 0.155 = square inches. 

Square meters X 10.764 = square feet. 

Cubic centimeters X 0.06102 = cubic inches. 

Cubic meters X 35.314 = cubic feet. 

Kilograms X 2.2046 = pounds. 


APPENDIX. 


493 


Metric-American Units Conversion Tables—( Continued .) 

Litres X 0.2642 = U. S. gallons. 

Calories X 3-968 = British thermal units. 

Chevaux X 0.9863 = horse power. 

Kilogrammeters X 7-233 = foot-pounds. 

Meters per second X 196.86 = feet per minute. 

Kilometers per hour X 0.6214 = miles per hour. 

Kilograms per square centimeter X 14.223 = pounds per square 
inch. 

Kilograms per litre X 8.347 = pounds per gallon. 

Calories per kilogram X 1.8= British thermal units per pound.. 
Calories per litre X 15.02 = British thermal units per U. S_ 
gallon. 

Kilometers per litre X 2.353 = miles per gallon. 

Imperial gallons X 1.2 = U. S. gallons. 

Kilowatts X 1-3405 = horse powers. 

(Centigrade temp. X 1.8) + 32 = Fahrenheit temp. 



INDEX. 


Page. 

A. L. A. M. Rating Formuia. 364 

Acceleration Forces, Table of. 46 

Adiabatic Changes of State. 15 

Adjustable Couplings. 268 

Admission Stroke. 19 

Air, Atmospheric . 6 

Air Cooled Exhaust Valves. 405 

Air Cooling More Difficult in Large Cylinders. 398 

Air Cooling of Four Cylinder Motors. 402 

Air Cooling of Two Stroke Motors. 430 

Air Cooling Used by Daimler. 398 

Air Displacement Constant Level System. 297 

Air Regulator, Hydraulic. 360 

Air Required to Burn Gasoline. 7 

Air, Weight and Properties of. 6 

Aluminum Alloys, Properties of. 271 

Amplitude of Vibration. 63 

Automobile Club of France Muffler Trials. 330 

Auxiliary Exhaust Ports. 254 

Babbitt Bushings, Manufacture of Split...'. 314 

Balance Weights, Calculation of. 62 

Ball Bearings, Radial. 194 

Base Circle Diameter. 235 

Baume Scale. 3 

Baush Multiple Drill. 307 

Bearing Bushings. 310 

Bearing Bushings, Relieving of. 315 

Bearing Cap Lock. 285 

Block Cylinder Castings. 73 

Block Cylinders. 91 

Blower, The. 403 

Blower Cooling. 402 

Bore and Stroke Ratio. 365 

Brake, The Motor as a. 374 

Breather Tubes. 281 

Breech Block Valve Closure, Winton. 105 

Bronze Bushings, Babbitt Lined. 313 

Brush Runabout, Balance Method of. 62 

Cadillac Sheet Metal Jacket. 90 

Cam and Roller Dimensions. 244 

Cam Followers, Types of... 228 

Cam for Mushroom Type Follower. 233 

Cam Gear Housing. 283 

Cam Grinding Attachment. 251 

Cam, Laying Out Acceleration. 230 

Cam Levers. 244 

Cam Opening Curve. 230 

Cam Outlines, Comparison of. 234 

Cam, Relation of Base Circle to Lift. 235 

Cam Side Thrust. 235 


495 





















































496 


INDEX. 


Page. 

Cams, Manufacture of. 249 

Cams, Offset. 237 

Camshaft Bearings. 248 

Camshaft, Dimensions of. 247 

Camshaft Drive, Arrangement of. 260 

Camshaft Drive, Silent Chain . 265 

Camshaft Gears, Dimensions of. 264 

Camshaft Gears, Methods of Securing. 264 

Camshafts, Integral and Built Up..,. 246 

Camshafts, Oiling. 249 

Car Speed Fluctuations in. 344 

Carbon Deposits Cause of Preignition. 399 

Carburetor Flanges. 322 

Carburetors for Two Stroke Motors. 422 

Chadwick Sheet Metal Jackets. 90 

Characteristic Curves. 369 

Charge Deflectors, Two Stroke. 413 

Charles’ Law. 12 

Cincinnati Vertical Miller.... 307 

Circulation Indicator. 303 

Columbia Lubrication System. 295 

Combustion Chamber, Spherical. 70 

Combustion Chamber Wall, Finished..... 399 

Compression Circulation. 24 

Compression Chamber Finishing Tool. 102 

Compression Leaks. 466 

Compression in Air Cooled Motor. 400 

Compression Relief Cams...... 253 

Compression Space Calculation. 78 

Compression Stroke. 20 

Compression, Work of. 25 

Connecting Rod Head. 208 

Connecting Rod, Cross Section of. 198 

Connecting Rod Head Oil Spoon. 213 

Connecting Rod Head, Size of Studs for. 211 

Connecting Rod, Hinged Cap. 208 

Connecting Rod, Crank Pin Bearings for. 210 

Connecting Rod, Length of. 197 

Connecting Rod Materials. 197 

Connecting Rod, Parts of. 197 

Connecting Rod, Variation of Cross Section of,» . 202 

Connecting Rod, Whipping Effect in. 203 

Connecting Rods, Machining of. 214 

Connecting Rods, Manufacture of . 214 

Connecting Rods, Offset...... 212 

Constant Level Stand Pipe. 302 

Cooling Flanges. 399 

Cooling Flanges, Form and Dimensions of. 400 

Counter Weights. 59 

Couplings, Magneto and Pump. 266 

Cradle Dynamometer, Electric. 455 

Crank Case, Barrel Type of. 276 

Crank Case, Bearings Between Halves of. 273 
























































INDEX. 497 

Page. 

Crank Case Bearing Caps. 285 

Crank Case, Bearings in Lower Half of. 274 

Crank Case, General Design of.. . 277 

Crank Chamber Design. 410 

Constant Acceleration Cam, Construction of. 230 

Continuous Web Construction. 289 

Convection of Heat from Air Cooled Cylinders... 400 

Cote Two Stroke Motor. 424 

Couplings, Magneto and Pump. 266 

Crank Case, in Upper Half of. 275 

Crank Case Inlets. 409 

Crank Case Materials.271 

Crank Case, Objects of. 271 

Crank Case, One and Two Cylinder.. 272 

Crank Case Sections, Thickness of. 279 

Crank Case Supporting Arms. 280 

Crank Case, Separate Supporting Arms. 281 

Crank Cases, Boring of. 307 

Crank Cases, Four and Six Cylinder. 273 

Crank Cases, Manufacture of. 307 

Crank Chamber, Automatic Circulation in. 297 

Crank Chamber Design, Two Stroke. 410 

Crank Chamber Joints Gas Tight, Making. 412 

Crank Moment. » . 49 

Crank Moment, Crank Angle Factor of. 54 

Crank Moment, Variation of. 382 

Crank Pin Lubrication. 291 

Crank Pin Turning Fixture. 189 

Crank Pin Turning Lathe. 189 

Crank Pins, Turning of. 191 

Crankshaft Balance. 59 

Crankshafts, Balancing . 171 

Crankshaft Bearing Metal. 286 

Crankshaft Blanks. 188 

Crankshaft Grinder. 19 1 

Crankshaft Material. 167 

Crankshaft Material, Heat Treatment of. 167 

Crankshaft Oil Ring. 293 

Crankshaft, Stresses in Members of. 183 

Crankshafts, Balancing . U 1 

Crankshafts, Ball Bearing. 192 

Crankshafts, Double Throw.176 

Crankshafts, Four Throw, Three Bearing. 178 

Crankshafts, Four Throw, Five Bearing. 179 

Crankshafts, Front and Rear Ends of. 187 

Crankshafts, Hollow'. I 9 ' 1 

Crankshafts, Machining of. I ^8 

Crankshafts, Multicylinder. *79 

Crankshafts, Single Cylinder. f . *75 

Crankshafts, Six Throw, Four Bearing. 186 

Crankshafts, Testing the Balance of. J 94 

Crankshafts, Types of. x ^9 

Crankshafts With Four Throws Between Supports. 181 























































4C8 


INDEX. 


Page. 

Cylinder Baffle Plate. 2 9 ° 

Cylinder Bolts. 2 7& 

Cylinder Bore, Length of. 81 

Cylinder Boring. 99 

Cylinder Boring Machine, Beaman & Smith. 99 

Cylinder Casting Testing Fixture. 97 

Cylinder Casting Testing Standard. 98 

Cylinder Castings, Annealing of. I0 3 

Cylinder Check Valve. 378 

Cylinder Design, Single. 82 

Cylinder Design, Twin. 83 

Cylinder Flange. 78 

Cylinder Grinder, Brown & Sharpe. 103 

Cylinder Grinding. I0 3 

Cylinder Grouping. 73 

Cylinder Head Closures. 106 

Cylinder Head, Thickness of. 77 

Cylinder Heads, Separate. 88 

Cylinder Jacket, Sheet Mttal. 90 

Cylinder Jacket, Thickness of. 77 

Cylinder, Jackson. 90 

Cylinder, L-Head. 86 

Cylinder Material. 74 

Cylinder Molding. 95 

Cylinder Pattern Making. 95 

Cylinder Ports, Symmetrically Arranged. 421 

Cylinder Stresses. 75 

Cylinder, T-Head. 83 

Cylinder Wall Area. 70 

Cylinder Wall, Common. 83 

Cylinder Wall, Pressure Against. 116 

Cylinder Wall, Side Thrust on. 160 

Cylinder Wall, Side Pressure on. 137 

Cylinder Wall Temperature, Effect on Power Output of. 373 

Cylinder Wall Temperature, Effect on Thermal Efficiency. 373 

Cylinder Wall Temperature, Limits of. 399 

Cylinder Wall Thickness. 76 

Cylinders Must Be Cooled, Why. 398 

Cylinders, Number of. 1 . 37 

Cylinders, Offset. 155 

Cylinders, Valve-in-Head. 89 

Diagonal vs. Lap Jointed Piston Rings. 127 

Diagram, Photographing the. 447 

Diehl Electric Dynamometer. 457 

Differential Piston Two Stroke Motor. 423 

Distributor Valve Two Stroke Motor. 426 

Drive, Arrangement of. 260 

Dynamometer Installation, A. C. A. 457 

Dynamometers .. 448 

Eccentric vs. Concentric Rings. 119 

Eight-Cylinder Motor, Balancing. 68 

Electric Cradle Dynamometer. 455 

E-M-F Lubrication System. 297 
























































INDEX. 


499 


Page-, 

Elmore High Duty Two Stroke Motor. 428 

Error in Manographs, Sources of. 440 

European Data and Deductions. 362 

Exhaust and Suction Curves. 31 

Exhaust Manifold. 323 

Exhaust Ports, Auxiliary. 254 

Exhaust Pressure Reducing and Control Valve. 378 

Exhaust . 22 

Exhaust Valves, Air Cooled. 405 

Expansion, Ratio of. 369 

Expansion, Work of. 29 

Explosion Pressure, Maximum. 131 

Explosion Pressure, Normal...;. 175 

Pace Adjustment of Manograph... .. 440 

Fan Bracket. 284 

Fan Bracket Adjustment. 107 

Fan Bracket Support. 107 

Fan Drive. 269 

Fan Dynamometer. 450 

Fan Dynamometer Calibration Curve. 454 

Fan Dynamometer Formula. 454 

Fan Dynamometer, Tracy’s. 453 

Fan Wheels . 347 

Faroux Power Output Formula. 3bs 

Firing Order, Effect of. 324 

Flexibility, Effect of Length of Stroke on. 368 

Flexibility of Starting of Motor and Car, Effect on. 336 

Flywheel, Bursting Strength of. 343 

Flywheel Capacity Required. 340 

Flywheel, Cast Steel. 347 

Flywheel, Centrifugal Force on. 342 

Flywheel, Effect on Flexibility of. 336 

Flywheel, Effect on Motor Starting of. 336 

Flywheel, Energy Stored in. 337 

Flywheel, Fan Type. 247 

Flywheel Flanges. 188 

Flywheel, Object of the. 332 

Flywheel Rim, Web or Spokes. 346 

Flywheels, Balancing. 350 

Flywheels, Keying of. 346 

Flywheels, Machining of. 348 

Foot-Pound, The. 10 

Force Feed. 290 

Force Feed and Splash System, Combined. 294 

Force Feed Non-Splash System. 291 

Ford Motor. 89 

Four Cycling of Two Stroke Motors. 422 

Four Cylinder Motor, Balancing. 66 

Four Cylinder Two Stroke Motor. 426 

Fuel Consumption Tests. 461 

Fuel Efficiency Tests. 227 

Franklin Motor, Volumetric Efficiency Test on. 223 

Gas Feed Through Valve. 222 























































500 


INDEX. 


Page. 

Gas Pressure Derived from Cylinder. 378 

Gases, Imperfect. 11 

Gases, Perfect. 11 

Gasoline, Calorific Value of. 

Gasoline, Chemical Composition of. 2 

Gasoline Combustion, Temperature of. *7 

Gasoline, Density of. 3 

Gasoline, Density Tests of. 4 

Gasoline Mixtures, Range of Inflammability of. 9 

Gasoline, Products of Combustion of. 7 

Gasoline, Temperature of Combustion of. *7 

Gasoline, Vapor Density of. 5 

Gasoline, Vapor Tension of. 5 

Gasoline, Viscosity of. S 

Governor, Hydraulic Type. 360 

Governor, Marine Type. 359 

Governor Springs, Calculation of. 354 

Governor, Vertical Type. 358 

Governors, Centrifugal.. 353 

Governors, Hunting of. 356 

Governors, Types of Centrifugal. 356 

Gravity Feed. 290 

Gravity Feed Non-Splash System. 294 

Guldner, Hugo, On Piston Rings. 120 

Hand Hole Cover. 287 

Hand Hole Plates.84-86 

Heat Absorbed by Cooling System, Amount of. 397 

Heat Absorption, Reduction of. 399 

Heat Discharged With Exhaust..197 

Heat Expansion, Allowance for. 148 

Heat Expansion, Allowance for Difference in. 109 

Heat Value, Higher. 8 

Heat Value, Lower. 8 

Heptane, Properties of. 3 

Hexane, Properties of. 3 

Holes and Drain Plugs, Filling. 305 

Hookham Joint. 267 

Horizontal vs. Vertical Cylinders. 69 

Horse Power Formula. 361 

Hospitalier Manograph. 435 

Hydraulic Dynamometers. 458 

Ignition and Combustion. 21 

Ignition and Combustion by Manograph, Studying. 441 

Ignition Points in Two Stroke Motors. 422 

Indicator Diagram, Standard. 139 

Indicators, Piston Type. 435 

Inertia Force, Crank Axle Factor of. 49 

Inertia of Gases, Effect of. 366 

Inlet Manifold. 320 

Inlet Valve, Automatic. 218 

Inlet Valves, Ported. 257 

Inspection Hole Covers. 288 

Isothermal Changes of State. it 
























































INDEX. 


50 i 


Page. 

Joule’s Equivalent. I0 

Knight Motor, The. ^gj 

Knight Motor, Test Results of. 39I 

Knight Sleeve Valve Motor. 6 9 

Knocks, Location of. 465 

Knox Cylinders. gg 

Late Ignition Shown by Diagrams. 446 

Legros Two Stroke Motor.426, 429 

Length of Stroke and Dead Centre Positions. 157 

Lubrication of Two Stroke Motors. 422 

Magneto Bracket. 287 

Manganese Bronze, Properties of. 271 

Manifold Joints. 321 

Manifolds, Integral. 94 

Manograph, The. 435 

Manograph, Construction of . 436 

Manograph Diagram. 443 

Manograph, Four-Cylinder . 442 

Manograph, Use of. 443 

Manographs, Quadruple. 442 

Marmon Oiling System. 293 

Mead Rotary Valve Motor. 395 

Mean Effective Pressure. 363 

Mechanical Equivalent of Heat. io 

Mechanical Oilers. 29 < 

Motor Tests . 433 

Motors, Weight of. 379 

Motor, The, as a Brake.374 

Muffler, Back Pressure of. 330 

Muffler, Baffle Plate Type. 328 

Muffler, Concentric Cylinder Type. 327 

Muffler Design. 326 

Muffler, Ejector Type. 328 

Muffler Location . 330 ■ 

Muffler Outlet . 330 

Muffler, Small Tube Type. 328 

Offset Engines, Cylinder Wall Thrust in. 160 

Offset Engines, Dead Centre Positions in. 157 

Offset Engines, Length of Stroke of. 157 

Offset Engines, Piston Acceleration in. 158 

Offset Engines, Piston Speed in. 158 

Offsetting, Effect on Balance of. 164 

Offsetting, Effect on Timing of. 164 

Offsetting, Object of. 155 . 

Offsetting, Reduction of Side Thrust by. 163 

Oil Conduits, Size of. 305. 

Oil Distributor. 304 

Oil Drain Plug.302, 305 

Oil Filling Holes. 305 

Oil Grooves in Piston Wall. 297 

Oil Guards. 286 

Oil Level Gauge. 303 

Oil Pump and Timer Drive. 268. 























































INDEX. 


502 

Page. 

Oil Pump, Gear Type. 2 98 

Oil Pump, Plunger Type. 3 QI 

Oil Pump, Sliding Vane Type. 3°o 

Oil Splash Trough, Adjustable. 39 1 

•Oil Strainers. 302 

Oil with Gasoline, Mixing. 2 9 % 

Otto Cycle. 

Otto Cycle, Application of Formula. 35 

Otto Cycle, Fundamental Relations. 38 

Otto Cycle, Piston Motion. 39 

Otto, Dr. N. A. 1 9 

Overlapping of Exhaust in Six Cylinder Motor. 325 

Packing Ring Arrangement. 383 

Panhard Knight Motor. 382 

Panhard-Levassor Sheet Metal Jackets. 90 

Pattern Making and Molding. 95 

Pentane, Properties of. 3 

Perfect Gases . 11 

Phase Adjustment of Manograph. 440 

Pierce-Arrow Lubricating System. 294 

Pin, Best Location of . 145 

Piston Acceleration. 43 

Piston Bearing Pressure. 142 

Piston Bosses. 134 

Piston Castings, Annealing of. 150 

Piston Castings, Cleaning. 150 

Piston Design, Sample. 149 

Piston Friction, Effect of. 144 

Piston Grinding Arbor. 144 

Piston, Hydraulically Forged Steel. 374 

Piston Pin, The. 

Piston Pin Bearing. 208 

Piston Pin Bearing, Adjustable. 208 

Piston Pin Bearing Bushing. 208 

Piston Pin, Best Location of. I4 5 

Piston Pin, Calculation of. X3 2 

Piston Pin Fastened in Connecting Rod. X36 

Piston Pin, General Design of. 130 

Piston Pin Lubrication. 293 

Piston Pins, Locking. 134 

Piston Position. 40 

Piston Pressure, Components of. 49 

Piston, Resultant Pressure on. 46 

Piston Ring Blanks. x 2I 

Piston Ring Blanks, Dimensions of. 126 

Piston Ring Clamp. I2 4 

Piston Ring Cutting-Off Tool. 122 

Piston Ring Grinders. 123 

Piston Ring Leakage Path. Iri 

Piston Ring Material for. II2 

Piston Ring Pliers. I30 

Piston Rings, Compound. I2 g 

Piston Rings, Eccentric vs. Concentric . no 























































INDEX. 503 

Page. 

Piston Rings, Fixture for Turning Eccentric. 121 

Piston Rings, Form of. II0 . 

Piston Rings, Hand Fitting of. I2 5 

Piston Rings, Handling. I3 o 

Piston Rings, Leakage Around. II0 

Piston Rings, Number and Location of. 148 

Piston Rings, Object of. io 8 

Piston Rings, Peening of. 126 

Piston Rings, Pinning of. 129 

Piston Rings, Radial Pressure of. 116 

Piston Rings, Stresses in. II3 

Piston Rings, Width of. 120 

Piston Side Thrust. 50 

Piston Side Thrust, Distribution of. 145 

Piston Speed . 41 

Piston Speed at Maximum Horse Power. 363 

Piston Speed and Acceleration. 158 

Piston Valves. 392 

Piston Wall, Thickness of. 147 

Pistons, Drill Work on. 153 

Pistons, Grinding of. 153 

Pistons, Lathe Work on. 150 

Pistons, Manufacture of. 150 

Plunger Pump . 301 

Port Capacity, Calculation of. 415 

Port Dimensions . 383 

Port Opening Area. 388 

Port Opening Integrals. 419 

Port Sizes . 414 

Power Indicator, The. 433 

Power from Given Cylinder Volume. 372 

Power Output . 361 

Power Output of Two Stroke Motor. 428 

Power Output, European Data on. 362 

Power Tests of Motor. 462 

Preignition Shown by Diagram. 446 

Pressure Variation in Cylinder. 23 

Proportion of Oil to Gasoline Used. 423 

Prony Brake, The. 448 

Pump Bracket. 287 

Pump Feed Splash System. 297 

Push Rod Guides, Yoke Retainers for. 24 

Push Rods, Adjustment in. 24 

Push Rods, Design of. 24 

Radiating Surface in Air Cooled Motor. 4 ° 1 ' 

Reciprocating Masses, Balancing. 60 

Reciprocating Parts, Weight of.44. 37 

Renault Motor, Valve Timing Tests on. 22\ 

Reynolds Rotary Valve Motor. 391 

Rings, Manufacture of.. 121 

Rings, Material of. 112 

Rings, Grinding . 123 

Rings, Hand Filing of... 125. 























































504 


INDEX. 


Page. 

Rings, Peening . l2 ^ 

Ring Blanks, Dimensions of. I2 ^ 

Rings, Special . I2 9 

Rings, Pinning . I2 9 

Rings, Number and Location of. *48 

Rochas. Beau .. 19 

Rope Brake, The. 449 

Rotary Valves. 393 

Rotating Parts, Balancing. 60 

“Running In” of Motor. 461 

Sample Design . *49 

Service of Motor Should Govern Ratio. 367 

Shims for Bearing Adjustment. 286 

Side Thrust Diagrams. *4° 

Side Thrust, Distribution of . 145 

Single Cylinder Motor, Balancing. 61 

Six Cylinder Motor, Balancing. 68 

Sleeve Valves, Lubrication and Cooling of. 39o 

Sleeve Valve Motion and Timing. 383 

Sleeve Valve Motor Packing Ring Arrangement. 383 

Sleeve Valve Motor Port Dimensions. 383 

Sleeve Valves, Power Consumed by. 390 

Sleeve Valves, Silent Operation of. 39° 

Slow Combustion Shown by Diagram. 446 

Spacing of Cylinders. 179 

Specific Heats. 12 

Specific Heats, Changes with Temperature of. 18 

Specific Power Output, Maximum. 372 

Speed Control. 352 

Splash Lubrication. 289 

Split Bushings, Mandrel for Turning. 313 

Split Bushings, Manufacture of. 310 

Spring Rests . 241 

Standard Spark Plug Shell, S. A. E. 473 

Standard Spark Plug Shell, De Dion (European). 473 

Standard Screws and Nuts, S. A. E. 479 

Standard Metric Threads. 480 

Starting Cranks. 316 

Starting Crank, Angular Position of. 317 

Starting Crank Automatic Lock. 3 I9 

Starting Crank Bracket. 3! 9 

Starting Crank, Effort Required on. 3!7 

Starting Crank Retainer. 284 

Starting Crank Support. 284 

Stearns-Knight Motor. 3 g 7 

Stethoscope . 4 66 

Supporting Arms, Separate. 281 

Supports, Four Throws Between. x g x 

Taper Pins . 4 _ 2 

T-Head Motor. _, 2 

Tangential Cam, Construction of. 229 

Temperature Scale, Absolute. Ix 

Test of Two Stroke Motor... . 























































INDEX. 


505 


Page. 

Testing Stands. 46 j 

Testing Castings . 97 

Thermal Efficiency. 3? 

Thermal Efficiency, Effect of Length of Stroke on. 365 

Thermal Unit, British. 10 

Thermodynamics, First Law of. 10 

Thermodynamics, Second Law of. 10 

Throttling of Two Stroke Motors. 4 >2 

Timing, Effect on. 164 

Timing Order, Effect of. 324 

Timing Marks on Flywheel. 464 

Tony-Huber-Peugeot Two Stroke Motor. 429 

Toothed Gearing . 474 

Torque Reaction . 63 

Traction Resistance . 469 

Transfer Passage Gauze Screen. 414 

Transfer Passage, Two Stroke. 4x3 

Twin Cylinders . 83 

Twin Cylinders, L-Head. 86 

Two Cylinder Motor, Balancing. 64 

Two Point Ignition, Effect of. 372 

Two Stroke Cycle, The. 407 

Two Stroke Motors, Constructional Details of. 431 

Two Stroke Port Sizes. 414 

Two Stroke Port Timing. 417 

Two Stroke Ports, Finishing of. 421 

V-Type Motor, Balancing . 65 

Valve Acceleration .... . . 234 

Valve Action, Silence of. 253 

Valve Cages . 71 

Valve Gear, Single Cam. 256 

Valve Guides . 224 

Valve Head, Form and Dimensions of. 221 

Valve-in-Head Motor, Power of. 72 

Valve Lift. 220 

Valve Material. 221 

Valve Mechanism, Clearance in. 224 

Valve Motion and Timing. 383 

Valve Opening Area. 219 

Valve Opening Integral. 234 

Valve Plug Yoke. 105 

Valve Plugs. 105 

Valve Seat, Angle of. 219 

Valve Spring Pressure Required. 238 

Valve Spring Rests. 241 

Valve Springs, Calculation of. 239 

Valve Timing. 225 

Valve Timing, Defects in.'• 464 

Valve Timing Tests. 226 

Valve Tool Lugs. 84 

Valves, Arrangement of . 69 

Valves, Concentric . 254 

Valves, Enclosed. 258 


























































506 


INDEX. 


Page. 

Valves, Mean Speed of Gas Through. 222 

Valves, Overhead. 245 

Valves with Cast Iron Heads. 222 

Vibration, Amplitude of. 63 

Vibration, Causes of.59 

Volumetric Efficiency. .... 223 

Volumetric Efficiency, Effect of Length of Stroke on. 366 

Water Jackets, Sheet Metal. 90 

Watson, Dr. W., Experiments in Motor Braking. 375 

Web Construction, Continuous. 289 

Weight of Reciprocating Parts. 44 

Weight Efficiency, Effect of Length of Stroke on. 367 

White & Poppe Power Output Experiments. 363 

Width of Ring. 120 


.jt 
















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PLATE I. 

Four Cylinder Truck Motor Built by the Pierce-Arrow Motor Car Co., Buffalo, N. Y. 


Cylinder Dimensions— 4 % x 6 Inches. 


















































































































































































































































































































































































































































































































































































































Four Cylinder Block Motor Manufactured by the F. I. A. T. Automobile Works, of Turin, Italy, and Poughkeepsie, N. Y. 

Cylinder Dimensions—110 x 150 Millimetres. 























































































































































































































































































PLATE III. 

Four Cylinder Motor Manufactured by the Nordyke & Marmon Co., Indianapolis, Ind 


Cylinder Dimension*—4^ xS Inches. 









































































































































































































































































































































































































































































































































































PLATE IV. 

Four Cylinder Block Motor Manufactured by Benz & Cie., of Mannheim, Germany 

Cylinder Dimensions—90 x 140 Millimeters. (Cylinders Are Offset 15 mm.) 






















































































































































































































































































































































































































































































































































Four Cylinder Motor Manufactured by the Reo Motor Car Co., Lansing, Mich. 


Cylinder Dimensions-4 x4^ Inches. (Cylinders Are Offset Inch.) 






















































































































































































































































































































































































































































































































































































































PLATE VI. 

Four Cylinder Block Motor Manufactured by Fabrique Nationale d’Armes de Guerre of Herstal, Belgium 

Cylinder Dimensions—80 x 120 Millimetres. (Cylinders Are Offset 25 mm.) 






















































































































































































































































































































































































































































































































































PLATE VII. 

Four Cylinder Motor Manufactured by the Velie Motor Vehicle Co., Moline, Ill. 

Cylinder Dimensions, 4 % x 5^6 Inches. 






























































































































































































































































































































































































































































































































PLATE VIII. 


Four Cylinder Block Motor, Manufactured by the Sunbeam Motor Car Co., Ltd., of Wolverhampton, England. 

Cylinder Dimensions—80 x 150 Millimetres. 



























































































































































































































































































































































































































































































































































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PLATE IX. 


Six Cylinder Motor Manufactured by the Continental Motor Manufacturing Co., Detroit, Mich. 


Cylinder Dimensions, 3 % x Inches. 


















































































































































































































































































































































































































































































































































































































































































PLATE X. 

Four Cylinder Piston Valve Motor Manufactured by the Model Gas Engine Co., Peru, Ind. 

Cylinder Dimensions, 3 94 x 5 94 Inches. 



















































































































































































































































































































































































































































PLATE XI. 

Six Cylinder Motor Manufactured by the Chalmers Motor Co., Detroit, Mich. 

Cylinder Dimensions, 4 x 55^ Inches. 







































































































































































































































































































































































































































































































PLATE XII. 

Four Cylinder Single Sleeve Valve Motor Manufactured by the Argylls, Ltd., of Alexandria, near Glasgow, Scotland. 

Cylinder Dimensions—101 x 130 Millimetres. 
















































































































































































































































































































































































































































































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PLATE XIII. 

Four Cylinder Power Plant Including Knight Type Sleeve Valve Motor Manufactured by the Moline Automobile Co., East Moline, Ills. 

Cylinder Dimensions, 4x6 Inches. 


































































































































































































































































































































































































































































































PLATE XIV. 

Six Cylinder Weidely Motor Manufactured by the Premier Motor Mfg. Co., Indianapolis, Ind. 

Cylinder Dimensions, 3% x 5^ Inches. 
































































































































































































































































































































































































































































PLATE XV. 

Four Cylinder Motor Manufactured by the Lozier Motor Co., Detroit, Mich. 

Cylinder Dimensions, 4*4 *6}^ Inches. 











































































































































































































































































































































































































































































■ — 











































































PLATE XVI. 

Six Cylinder Air Cooled Motor Manufactured by the H. H. Franklin Mfg. Co., Syracuse, N. Y. 

Cylinder Dimensions—4 x 4 Inches. 








































































































































































































































































































































































































































































































































































































































































































































































































































































PLATE XVII. 

Six Cylinder Motor and Clutch Manufactured by the Northway Motor and Mfg. Co., Detroit, Mich. 


Cylinder Dimensions, 4 % x 4 94 Inches. 







































































































































































































































































































































































































































































































































































































































































































































































PLATE XVIII. 

Four Cylinder Motor Built by the Rutenber Motor Co., Marion, Ind. 


Cylinder Dimensions, 4 la x Inches. 











































































































































































































































































































































































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PLATE XIX. 

Four Cylinder Power Plant Manufactured by the Paige-Detroit Motor Car Co., Detroit, Mich. 

Cylinder Dimensions, 4x5 Inches. 




























































































































































































































































































































































































































































































































































































































































































































































































































PLATE XX. 

Four Cylinder Stellite Motor Manufactured by the Wolseley Motor Car Co., Birmingham, England. 

Cylinder Dimensions, 62 x 89 mm. 



















































































































































































































































































































































































PLATE XXI 

Double Cylinder Opposed Motor formerly Manufactured by the Maxwell-Briscoe Motor Co., of Tarrytown, N. Y. 

Cylinder Dimensions-41& x4 Inches. 














































































































































































































































































































































































































































PLATE XXII 

Eight Cylinder V Type Motor Manufactured by the Ferro Machine and Foundry Co., Cleveland, O. 

Cylinder Dimensions, 3 ^4 x 4 Inches. 





























































































































































































































































































































































































































































































































































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PLATE XXIII. —Chart for Determining Piston Displacement 








































































































































































































































































































































































































































































































































































































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